design and internal flow analysis of a ducted contra
Post on 20-Oct-2021
0 Views
Preview:
TRANSCRIPT
DESIGN AND INTERNAL FLOW ANALYSIS OF A DUCTED CONTRA-ROTATING AXIAL FLOW FAN
Ali Mohammadi Amirkabir University of Technology
Tehran, Iran
Masoud Boroomand Amirkabir University of Technology
Tehran, Iran
ABSTRACT This paper presents the design procedure of a ducted contra-
rotating axial flow fan and investigates the flow behavior inside it
using ANSYS CFX-15 flow solver. This study investigates
parameters such as pressure ratio, inlet mass flow rate and
efficiency in different operating points. This system consists of
two rotors with an outer diameter of 434 mm and an inner
diameter of 260 mm which rotate contrary to each other with
independent nominal rotational speeds of 1300 rpm. Blades’
maximum thickness and rotational speeds of each rotor will be
altered as well as the axial distance between the two rotors to
investigate their effect on the overall performance of the system.
Designed to deliver a total pressure ratio of 1.005 and a mass
flow rate of 1.8 kg/s at nominal rotational speeds, this system
proves to be much more efficient compared to the conventional
rotor-stator fans. NACA-65 airfoils are used in this analysis
with the necessary adjustments at each section. Inverse design
method is used for the first rotor and geometrical constraints are
employed for the second one to have an axial inlet and outlet
flow without using any inlet or outlet guide vanes. Using free
vortex swirl distribution method, characteristic parameters and
the necessary data for 3D generation of this model are obtained.
The appropriate grid is generated using ATM method in
ANSYS TurboGrid and the model is simulated in CFX-15 flow
solver by employing k-ε turbulence model in the steady state
condition. Both design algorithm and simulation analysis
confirm the high anticipated efficiency for this system. The
accuracy of the design algorithm will be explored and the most
optimum operating points in different rotational speed ratios
and axial distances will be identified. By altering the outlet
static pressure of the system, the characteristic map is obtained.
INTRODUCTION There is currently a growing tendency toward developing
contra-rotating fans and compressor stages due to their
compactness in length and reduction in weight which is the result
of omitting stators from the conventional models. Many
researchers working in this area have reported higher quantities in
efficiency and pressure ratio for contra-rotating fans, propellers
and compressor stages in comparison to rotor-stator assembly.
Started in 1930s, Lesley [2] showed that an addition of fixed
counter-propeller blades increases the efficiency of a four-blade
fixed counter-propeller in combination with a two-blade rotating
propeller by two percent. In 1980s, Sharma et al. [3] examined
the effect of altering rotational speed ratios of two rotors and also
different axial spacing between them in one stage of a compressor
which had blades with 0.66 hub-tip ratio. It was concluded that
alteration of these factors has strong influence on the stalling
behavior of the stage. Also it was mentioned that if the second
rotor contra-rotates 50% faster than the first one, rotating stall
phenomenon would be suppressed and in case of large axial
distances, contra-rotation would lose its benefits. In 2009, an
increase in efficiency and a decrease in fuel consumption rate
were reported by Min et al. [4] in a contra rotating propeller
system. They claimed that by using the rotational flow energy
behind the first rotor, the propulsive efficiency of the system
would be improved. Later, Nouri et al. [5] further analyzed the
effect of the rotational speeds ratios of the rotors and the axial
distance between them in a low number of blade contra-rotating
axial flow fan. They reported optimum quantities for these
parameters and validated the results experimentally. Finally,
Mistry et al. [6] conducted a similar experiment on high aspect
ratio contra-rotating axial fan stage. They reported high quantities
of efficiency in design and off design operating points. In this
research, after choosing an optimum blade thickness, operation of
a high number of blades axial flow fan in different axial distances
and also with different rotational speed ratios would be
investigated to indicate its characteristic map. In near future, an
actual model will be built based on the design procedure
introduced in this article.
Proceedings of the ASME 2014 International Mechanical Engineering Congress and Exposition IMECE2014
November 14-20, 2014, Montreal, Quebec, Canada
IMECE2014-39883
1 Copyright © 2014 by ASME
NOMENCLATURE = Flow Angle in the Absolute Reference
= Flow Angle in the Relative Reference
' = Blade’s Angle
= Camber Angle
i = Incidence Angle
= Deviation Angle
= Stagger Angle
= Loading Coefficient
= Flow Coefficient
= Heat Capacity Ratio
Cp = Molar Specific Heat at Constant Pressure = Solidity = Density
Re = Reaction Ratio
T = Temperature
P = Pressure
N = Rotational Speed
R = Radius
RR = Rotational Speed Ratio
U = Linear Speed
DF = Diffusion Factor
R.P. = Required Power
L = Chord’s Axial Length
h = Blade’s Height
s = Axial Spacing
NoB = Number of Blades
A.R. = Aspect Ratio
A = Area
D = Diameter
P.R. = Pressure Ratio = Efficiency
m = Mechanical Efficiency (Engine)
= Loss Coefficient
= Loss Coefficient
m = Mass Flow Rate
Cx = Axial Flow Speed
w = Tangential Velocity
Cl = Lift Coefficient
Cd1 = Initial Drag Coefficient
Cda = Annulus Drag Coefficient
Cds = Secondary Drag Coefficient
tb/c = Thickness
Kt,i = Incidence Thickness Correction Factor
Kt,δ = Deviation Thickness Correction Factor
Ksh = Shape Factor
Subscripts:
1 = First Rotor
2 = Second Rotor
11 = Inlet of First Rotor
12 = Outlet of First Rotor
21 = Inlet of Second Rotor
22 = Outlet of Second Rotor
0 = Zero Camber Condition
10 = Parameter for 10% Thick Profile
o = Total Thermodynamic Condition
h = Blade’s Hub
m = Blade’s Mean-line
t = Blade’s Tip
DESIGN OF ROTORS Based on the fundamental concepts in Aungier [1], an inverse
design procedure is developed for the first rotor. The second rotor
is designed using some geometrical constraints between the outlet
flow angles of the first rotor and the inlet flow angles of the
second one. Simplified Radial Equilibrium is employed for this
procedure. Two 750 watt Electrical Motors with a maximum
rotational speed of 1385 rpm are assumed to be used in future
experimental analysis. Thus, the required power in the operating
points must not exceed this value. Setting the hub diameter (Dh) in
accordance with this engine diameter to be 26 cm and also
employing a hub-tip ratio of 0.598, the tip diameter (Dt) of these
contra-rotating rotors would be 43.5 cm. A 2.5 mm tip clearance
length is anticipated at the shroud section for safety consideration
in future experimental studies. Thus, the casing diameter (Dc)
would be 44 cm.
To have an axial inlet and outlet flow without using any Inlet
and Outlet Guide Vanes, the corresponding flow angles in the
absolute frame is set to be zero (Equation 1). According to the
selected electro-motor specifications, nominal design rotational
speeds are set to be 1300 rpm for both rotors. It is desired to have
an inlet mass flow rate of 1.8 kg and a total pressure ratio of
1.005. The mechanical efficiency of the selected electro-motor
(ηm) is 70% and the isentropic efficiency of the system (η) is
assumed to be 85%. Standard atmospheric condition is employed
(To = 288 K and Po = 101.3 kpa) which gives a density value of
1.226 kg/m3 in combination with equation 3 in an iterative
procedure. Equation 4 gives us the linear speed at each section
and equations 5 and 6 give us the required power for the first
rotor of this contra-rotating fan in the design rotational speed
which is equal to 624.326 watt considering the electro-motor’s
mechanical efficiency. Using equations 7 to 9, some
dimensionless variables such as the loading coefficient, the flow
coefficient and the reaction ratio are obtained. Table 1
summarizes the initial quantities obtained from equations 1 to 9.
01211 (1)
)(4
22
ht DDA
(2)
A
mCx
(3)
2 Copyright © 2014 by ASME
)(2 NRU (4)
]1..[
1
11
RPTT o
o
(5)
m
op TCmPR
2..
1
1
(6)
2
1
U
CT po
(7)
U
Cx
(8)
11tan2
1Re
(9)
Table 1 - Initial Parameters
Parameter Quantity
A (m2) 0.099
Cx (m/s) 14.835
Um (m/s) 23.596
ΔTo1(K) 0.241
R.P.1 (watt) 624.326
m1 0.436
m1 0.628
Re1m 0.782
As well as using the obtained dimensionless variables,
some geometrical constraints are employed between the first
rotor’s outlet flow angles and the second rotor’s inlet flow
angles in the absolute and also in the relative reference systems
to obtain other flow angles. Equation 11 states that the exiting
flow from the first rotor must enter the second rotor with a same
flow angle. The minus sign indicates the change of coordinate
system which is due to the contra-rotation of two rotors. The
velocity triangles in Figure 1 show the relation between these
flow angles.
)tan)Re1(2
(tan 11
1
11
12 m
m
mm
(10)
mm 2112 (11)
)tan1Re2
(tan 11
1
11
12 m
m
mm
(12)
)tanRe2
(tan 12
1
11
11 m
m
mm
(13)
)tan
(tan 12211
21
x
mxmmm
C
CUU
(14)
)(tan 21
22
x
mm
C
U
(15)
.
Figure 1 - Velocity Triangles of Rotors
To obtain the optimum solidity of blades for each rotor, the
procedure presented in [7] is employed as follows:
n
opt BA 2 (16)
where,
)(042231.00197.0 21 A (17)
))}(002368.033303.0(
)(427.13exp{
21
21
B
(18)
)(04677.08592.2 21 n (19)
Choosing a desirable aspect ratio of 2 for each rotor, the
number of blades and also the axial distance between them can
be determined using equations 20 to 22. Quantities obtained for
the number of blades are rounded to the nearest odd number.
Therefore, the first rotor will consist of 27 blades and the
second one consists of 21 blades. As the number of blades is
determined, axial spacing between them and also axial chord
lengths would be calculated again using equations 21 and 22.
Assuming the axial chord lengths to be constant in each section
and employing free vortex swirl distribution method, similar
calculation is done for all other sections of each rotor. Table 2
summarizes some of these obtained quantities. Then, equations
25 to 32 are used respectively to calculate the diffusion factor,
different drag coefficients, lift coefficient and loss coefficient.
Finally equation 33 gives the efficiency of the system. Further
explanations of these parameters are available in [1].
3 Copyright © 2014 by ASME
Table 2 - Geometrical Data of Both Rotors
hub
Rotor 1
Mid
Shroud
Hub
Rotor 2
Mid
Shroud
Inlet Relative Flow Angle (degrees) 50.012 57.581 63.298 64.721 66.273 68.533
Outlet Relative Flow Angle 15.012 41.156 55.097 50.027 57.581 63.298
Solidity 1.456 1.103 0.873 1.066 0.807 0.639
Incidence Angle 0.385 -0.792 -1.065 -2.688 -4.086 -5.678
Camber Angle 46.38 26.696 17.117 28.678 23.538 22.299
Deviation Angle 11.75 9.478 7.85 11.295 10.761 11.386
Stagger Angle 26.451 45.026 55.805 53.071 58.589 63.061
AR
hLm
(20)
mopt
mm
Ls
(21)
m
m
s
RNoB
2
(22)
tan xCw (23)
)tan(tan 21 xCw (24)
11
2
21
wL
sw
w
wDF
(25)
)tan(tan2
1tan 21 m
(26)
1
1
2cos
1w
(27)
mdL
sC 3
1 cos
(28)
mdm
l CL
sC
tan
)tan(tancos21
21
(29)
h
sCda 02.0
(30)
2018.0 lds CC (31)
dsdaddt CCCC 1 (32)
m
dtmt
CwL
3
1
cos
21 t
(33)
Aungier [1] mentioned that NACA-65 airfoil’s general
profile can be used to determine the coordinates of blades at
each desired section. This general set of coordinates is
available in Table 4-1 [1]. For this procedure, a zero camber
lift coefficient must be calculated. Using equations 34 to 45,
the incidence and the deviation angles are determined. Ksh is
set to be 1 for NACA-65 profiles [1]. As it will be introduced
further in the following section, maximum thickness of the
blades which is indicated by tb/c is assumed to be 0.08. This
means the first rotor blades are 3.528 mm thick and the second
rotor blades are 3.32 mm thick at their most.
q
bit ctK )/10(, (34)
])(1.0[
28.0
3.0
c
tq
b
(35)
]4/)70exp[(
1.0)3.2exp(465
)(
1
3110
*
0
p
i
(36)
160/914.0 3p (37)
43.05.1
)90/(06.0025.0
)2.11(
12n
(38)
210
*
0,
* )( niKKi itsh (39)
)09.167.1(
1
9.1
110
*
0
)90/(
]374.0[01.0)(
(40)
bmm /0.1 (41)
32
0.1 316.0132.0074.0249.0 xxxm (42)
385.017.09625.0 xxb (43)
2
, )/(5.37)/(25.6 ctctK bbt (44)
mKK tsh 10
*
0,
* )( (45)
To indicate the camber angles, equation 39 and 45 are
solved in an iterative manner. Alternatively, equations 46 to 49
may be employed directly by introducing A2 & B2 which are
4 Copyright © 2014 by ASME
derived from equations 34 to 45 and also from equations 50 to
53. Finally, equation 54 provides the zero camber lift
coefficient. Therefore, coordinates of blades at all sections are
now derived. Table 2 summarizes the obtained values at three
sections and Figure 1 summarizes the whole design process.
1
)()(
0.1
*
02
2
10
*
012
m
B
n
iA
(46)
1
)1(
0.1
0.1
2
22
m
m
n
nB
(47)
2
2'
11B
A
(48)
2
10
*
011
'
112
'
12
)()1(
n
in
(49)
'
12
'
111 (50)
'
11111 i (51)
'
12121 (52)
2
1'
111
(53)
4tan
1103.0
1 10
lC
(54)
Figure 2 - Summary of the Design Procedure
NUMERICAL MODEL Data obtained using the procedure explained in previous
section is used to generate the 3D model in ANSYS TurboGrid
which is also used for grid generation. ATM optimized
(Automatic Topology and Meshing) method is chosen over
other methods due to its better adoptability to blades surfaces.
Various grids with different number of elements are analyzed
for grid sensitivity analysis. The inlet total temperature and
pressure are set to be 288K and 101.3 kpa respectively while
the outlet boundary condition is determined using the outlet
static pressure. No slip wall condition is used for the hub and
blade sections and contra-rotating walls are used for the
shroud sections. Table 3 summarizes the results for grid
analysis while the outlet static pressure is 101.6 kpa. Also,
Figure 3 shows the generated model of the first rotor.
Table 3 - Grid Sensitivity Analysis for Contra-Rotating
Fan
1st Rotor
Surrounding
Nodes
2nd
Rotor
Surrounding
Nodes
Total
Pressure
Ratio
Total
Efficiency
52488 59236 1.0047 83.9877
129438 128840 1.0048 86.8611
241083 262742 1.0049 87.312
Figure 3 - Generated 3D Model for the First Rotor
The difference in total pressure ratio of the second and the
third sets of grid is less than 0.01% and the difference between
the efficiencies of these two sets is less than 0.52%. Thus, the
second set is selected to continue the analysis as using finer
grid shows no major improvement on the results. A gap equal
to 2.5 mm is anticipated between the blade’s tip and the
covering case to simulate the tip clearance length. Then the
model is simulated and solved using ANSYS CFX-15 flow
solver using standard atmospheric condition and also the
design parameters discussed before. k-ε turbulence model in
5 Copyright © 2014 by ASME
steady state condition is used as it is widely suggested in other
published researches in this area as in [6]. The final
computational grid consists of one single passage for each
rotor. The simulation is then generalized for the whole system
using periodic condition between the flow passages. The
second rotor’s domain starts exactly where the first domain
ends as it is shown in Figure 4. For cases with an axial distance
greater than 2 cm, a stationary part is added to the model to
increase the dignity of the solution. Figure 5 shows each of
these models.
Figure 4 - Generated Grid around the First Rotor (right)
and the Second Rotor (Left)
Figure 5 - Components in Different Axial Distances: a- 2cm
b- 4 to 12cm
For validation of the turbulence model, NASA Stage 37 [8]
was analysed in two different rotational speeds (70% and 90% of
the Nominal Rotational Speed). The reason that NASA stage 37
is preferred over other published models such as NASA rotor 67
for validation in this project is that this model is consisted of two
consecutive frames. Similarly, the current understudied project
consists of two frames in most cases. In 70% of nominal
rotational speed, the flow is mainly completely in the subsonic
flow regime; thus, it perfectly matches this subject. It should be
mentioned that data (geometry of the blades and simulation data)
for a more similar case which is consisted of two consecutive
rotors was not fully available to be used. Table 4 contains
information regarding the grid sensitivity analysis of this model
while operating with an outlet static pressure of 101.3 kpa. The
difference between the total pressure ratio in the second and the
third sets of generated grids is 0.015% and the difference
between the efficiencies is 0.064%. So the second set is chosen
for validation purposes as no major improvement is made in
results using finer grids.
Table 4 - Grid Sensitivity Analysis for NASA Stage 37
1st Rotor
Surrounding
Nodes
2nd
Rotor
Surrounding
Nodes
Total
Pressure
Ratio
Total
Efficiency
45710 37408 1.291 83.7326
169735 169101 1.2905 84.4847
323282 329529 1.2903 84.5393
Figure 6 shows a comparison between the experimental data
of NASA stage 37 report and the results obtained from
simulation of this model. Comparison of the pressure ratios in
both rotational speeds which is presented in Table 5 shows that
at its most, the results are less than 6% different which shows a
great agreement. As the validation of the turbulence model is of
interest here and the results show a similar trend and an
acceptable difference, k-ε will be used to continue the analysis.
Figure 6 - Comparison between the Simulation Results and
the Experimental Report of NASA Stage 37
6 Copyright © 2014 by ASME
Table 5 - Comparison of Results in Similar Inlet Mass Flow
Rates
Inlet Mass Flow Rate (kg/s) Difference in P.R. (%)
18.89 5.519
17.91 1.18
17.26 1.147
15.38 4.161
14.81 3.168
14.11 2.823
13.24 2.284
RESULTS AND DISCUSSION
Effect of Blade’s Thickness To investigate the effect of
blade’s thickness on the overall performance of the system, three
different quantities for blade thickness are chosen and used for
preliminary simulation which is presented in Table 6. For this
step, the outlet static pressure is chosen to be 101.65 kpa.
Table 6 – Effect of Various Airfoil Thicknesses
Max. Blade
Thickness (tb/c)
Inlet Flow
Rate (kg/s)
Pressure
Ratio
Efficiency
0.08 1.6778 1.0051 83.8157
0.1 1.6332 1.005 83.1111
0.12 1.5771 1.005 82.1774
Table 7 - Comparison of Parameters between Different
Blade Thicknesses and tb/c = 0.08
Max. Blade
Thickness
(tb/c)
M.F.R
Difference
% (kg/s)
P.R.
Difference
%
Efficiency
Difference
%
1 -2.658 -0.0099 -0.841
1.2 -6.001 -0.0099 -1.955
It is seen that thickening the blade profiles decreases the
possible inlet mass flow rate as it blocks some of the available
flow passage. Also, the total pressure ratios for the second and
third blade thicknesses are decreased compared to the one
related to maximum blade thickness of 0.08. The same pattern
occurs in the efficiency of the system. Table 7 shows the
comparison between the parameters obtained by the first blade
thickness (tb/c = 0.08) and the two other sets. Based on the
stated results, it is decided to choose 0.08 for the maximum
thickness chord ratio to further analyze the behavior of this
contra-rotating fan.
Accuracy of the Design Algorithm Results obtained from
simulation is used to verify the accuracy of the design algorithm.
Table 8 shows the consistency of results obtained from the
design algorithm and the simulation analysis. It is seen that the
total pressure ratio is slightly over predicted and the efficiency is
slightly under predicted in the design algorithm.
Table 8 - Validation of the Design Algorithm
Parameter Design
Algorithm
Simulation
Analysis
Difference
(%)
Inlet Flow Rate (kg/s) 1.8 1.8 0
Total P.R. 1.005 1.0046 0.039
Total Efficiency 85.0 86.4 1.647
Effect of Axial Distance Variation Six different axial
distances between two rotors are investigated starting from 2cm
(50% chord of the first rotor’s blade) to 12cm (300% chord of
the same blade). For the axial distance of 2cm, two rotating parts
are considered. For all other five cases from 4cm to 12cm, three
components are considered including one stationary part
between the rotors.
Figures 7a to 7c show the effect of variation in axial
distance on the operation of this contra-rotating system. Figure
7a shows that variation in axial distance has a negligible effect
on the inlet mass flow rate when the outlet static pressure is
altered. Figure 7b show that increasing this distance doesn’t
necessarily improve the pressure ratio of the system. Yet, Figure
7c shows that the CRF (Contra-Rotating Fan) operates with the
highest efficiency in the minimum axial distance. There can be
two reasons for this phenomenon. The first reason is that the
rotational energy produced by the first rotor can be best utilized
by the second rotor in case of minimum axial distance.
Increasing this distance results in dissipation of the rotational
energy behind the first rotor and thus decreases the efficiency of
the system. The second reason is that at this minimum axial
distance, the flow will be delivered to the second rotor at its
most precise quantity of the inlet flow angle. All other cases
from 4cm to 12cm show a very similar pattern in efficiency
while the inlet mass flow rate is altered.
Another interesting phenomenon is that when the system
is operating at the minimum anticipated length, the stall margin
would be widened. Figure 7a to 7c show that increasing this
distance from 2cm to 6cm limits the operating range. However,
between 6cm and 12cm, there seems to be no variation in this
issue. The maximum outlet static pressure which corresponds
to the stall limit is plotted vs. different axial distances in Figure
8. This can also be observed in relative Mach no. plots as well.
The velocity contours of the first rotor for all axial distances
are similar. Thus, only the one related to the minimum axial
distance is presented in Figure 9. Over the first rotor, the flow
is completely in low subsonic regime as anticipated.
7 Copyright © 2014 by ASME
Figure 7a-c Performance Parameters of CRF in Different
Axial Distances
Figure 8 - Stall Outlet Static Pressure vs. Axial Distance
Figure 9 - Relative Mach No. Contours for the First Rotor
in Axial Distance = 2cm
In the second rotor, there is a small change of relative
velocity in different axial distances. While operating at the
minimum distance, the blades experience a slightly less
relative Mach no. compared to other cases which is shown in
Figure 10a. Increasing the distance from 4cm to 12cm has no
major influence on this parameter. This is shown in Figures
10b and 10c where the Mach no. contours are similar. All
contours in Figure 9 and 10a to 10c are obtained when the
outlet static pressure is 101.6 kpa. Because of better efficiency
of the system in the minimum axial distance which is mentioned
above, this case is chosen to be used for investigation of the
effect of rotational speed ratio on this system.
Effect of Rotational Speed Ratio Introducing RR as the
rotational speed ratio, five different cases will be analysed.
1
2
N
NRR
(55)
8 Copyright © 2014 by ASME
Figures 10a-c Relative Mach No. Contours for the Second
Rotor in Axial Distances of a- 2cm, b- 4cm and c- 12cm
These five cases are: 1300 - 1150 (RR=0.884), 1300 -
1225 (RR=0.942), 1300 - 1300 (1.00), 1300 - 1385 (1.065)
and 1385 - 1300 (RR=0.938). Figure 10a to 10c show that
when N1 is fixed at 1300 rpm, increasing N2 improves the
performance of the system. Figure 11a shows that increasing
N2 results in increasing of the inlet mass flow rate while the
outlet static pressure is constant. Also Figures 11b and 11c
show that the system delivers higher pressure ratio and
operates with a better efficiency when N2 is increased. The
reason is the suction effect of the second rotor. Increasing N2
increases the suction effect of this rotor which eliminates the
wake produced by the first one. Comparing 1300 - 1225 set to
1385 - 1300 set which has nearly the same RR (0.94) shows
that the second set has better performance. This is because
increasing N1 to 1385rpm decreases the wake produced after
the first rotor and also the afterwards flow separation.
Figure 11a-b Performance Parameters of CRF in different
Rotational Speed Ratios
9 Copyright © 2014 by ASME
Figure 11c - Performance Parameters of CRF in different
Rotational Speed Ratios
Figures 12a to 12c show the velocity contours of the first
rotor operating in three different RR. It is seen that although
N2 has increased from 1150 rpm to 1385 rpm, there is no
major change in magnitude or direction of velocity vectors of
the first rotor while N1 is constant. Increasing N1 to 1385 rpm
results in higher magnitudes of velocity as it can be seen in
Figure 12c.
Figure 12a-c Velocity Contours for the First Rotor in
a- 1300-1150 set, b- 1300-1385 set and c- 1385-1300 set
Figures 13a to 13c show the velocity contours for the
second rotor. Comparing figure 13a and 13b shows that
increasing N2 has a little effect on the direction of the velocity
vectors. However it increases the velocity magnitude. On the
other hand, increasing N1 has a little effect on either the
direction and also the magnitude of velocity vectors of the
second rotor.
10 Copyright © 2014 by ASME
Figure 13a-c Velocity Contours for the Second Rotor in
a- 1300-1150 set, b- 1300-1385 set and c- 1385-1300 set
CONCLUSIONS A step by step design procedure of a contra-rotating axial
flow fan is presented. Then the 3D model is generated and the
appropriate grid is applied to the model in ANSYS Turbogird.
Finally the flow field around the rotors is simulated using CFX-
15 flow solver to determine characteristics of this system in
various operating conditions. Due to the results obtained, the
following conclusion can be mentioned:
1. Airfoil thickness has an important influence on the
efficiency and pressure ratio of the system. It is decided to
choose NACA-65 airfoil profiles with a maximum
thickness to chord ratio of 0.08 due to its higher inlet mass
flow rate in the same operating condition compared to
other cases and also because of its higher efficiency.
2. Analysis of different axial distance between the rotors
show that an axial distance equal to 50% of chord is the
most optimum distance for this CRF. It is shown that
employing this axial distance results in obtaining higher
efficiency and also a wider operating range which delays
the stalling condition. Further increase in the axial length
between two rotors, from 6cm to 12cm shows no
significant influence on the performance of the system.
3. Analysis of the rotational speed ratios (RR) indicates
that as the rotational speed ratio increases, better
performance is obtained from this system in terms of
efficiency and pressure ratio. This is because of the
stronger suction effect imposed by the second rotor on the
first one as N2 increases. Also, increasing N1 reduces the
produced wake by this rotor which results in better
performance. The best performance is achieved while the
system is operating with a RR ratio of 1.065 and an axial
distance equal to 50% of the first blade’s axial chord.
Based on the results obtained, an experimental test stand for
this contra-rotating axial flow fan will be built at DANA
Research Laboratory. As the experimental data is not available at
this time, a previous published case of NASA is used for
validation of the turbulence model and results of the simulation
analysis. Once the system is constructed, experimental data will
be compared to what obtained from simulation. Further
computational study is also underway to explore the effect of
different tip clearances and also higher rotational speed sets.
REFERENCES 1. Aungier, R.H. [2003], Axial Flow Compressors, A strategy
for Aerodynamic and Design Analysis Book, ASME Press.
2. Lesley, E. [1993], Experiments with a Counter-Propeller
Tech. Rep. 453, National Advisory Committee for
Aeronautics, Washington, USA.
3. Jain, Y. , Pundhir, D. and Sharma, P. [1998], A Study of
some Factors Affecting the Performance of a Counter-
Rotating Axial Compressor Stage, Proceedings of the
Institution of Mechanical Engineers Part A. Power and
Process Engineering, pp. 15-21, New Delhi, India.
4. Chang, B.-J., Min, K.S. and Seo, H.-W. [2009] Study on
the Contra-Rotating Propeller System Design and Full
Scale Performance Prediction Method, International
Journal of Naval Architecture and Ocean Engineering, pp.
29-38, Ulsan, Korea.
5. Bakir, F., Nouri, H., Ravelet, F. and Sarraf, C. [2011],
Experimental Investigations on the Design of a Ducted
Counter-Rotating Axial Fan System, 46th Symposium of
Applied Aerodynamics - Aerodynamics of Rotating Blades,
Orleans, France.
6. Mistry, C. and Pardeep, A. [2013], Effect of Variation in
Axial Spacing and Rotor Speed Combinations on the
Performance of a High Aspect Ratio Contra-Rotating Axial
Fan Stage, Proceedings of Institution of Mechnanical
Engineers, Part A: Journal of Power and Energy
7. Tournier, J. and El-Genk M. [2010], Axial Flow Multi-
Stage Turbine and Compressor Models, Journal of Energy
Conversion and Management, pp. 16-29
8. Moore, R. and Reid, L. [1978], NASA Technical Report
Paper 1337, Design and Overall Performance of Four
Highly Loaded, High-Speed Inlet Stages for an Advanced
High-Pressure Ratio Core Compressor
11 Copyright © 2014 by ASME
top related