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Transcript of Air Flow Patterns
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I C h e i T l E 0263-8762/04/S30.00 0.00
2004 Institution of Chemical Engineers
Trans
IChemE, Part A, October
2004
Chemical Engineering Research and Design, 82 A10): 1344—1352
IR FLOW
P TTERNS
IN DEHUMIDIFIER
WOOD DRYING KILNS
Z .
F. SUN
1
, C. G.
C A R R I N G T O N 1
J . A.
A N D E R S O N 2
and Q. SUN
1
1Physics
Department, University of Otago, Dunedin, New Zealand
^Energy
Group Ltd, Dunedin, New Zealand
B
y simulating the airflow
patterns,
velocity and
pressure
distributions in an industrial
dehumidifier wood drying
k i l n ,
it is shown that typical dehumidifier system configur
ations create a risk of high levels of air recirculation at the dehumidifier, wi th adverse
implications for dryer capacity and efficiency. The simulation results also show that, for high
efficiency,
it is important to avoid air recirculation. An alternative air flow configuration,
which
could
achieve
this result using a single set of
fans,
is
presented
and its
performance
assessed.
Keywords: wood drying;
heat
transfer; mass transfer; mathematical modelling;
heat
pump;
dehumidifier.
INTRODUCTION
Air f low
design is potentially important in the design of
dehumidifier drying kilns which
operate
as closed, fu l ly -
recirculatory
systems.
In this
paper,
we show how the
performance
of a dehumidifier wood drying ki ln can be
impaired by a mismatch between the ki ln airflow system
and the dehumidifier. In turn, the poor
performance
of
the dehumidifier reduces the overall efficiency of the
ki ln ,
resulting in
increased
drying time and energy use.
In industrial wood drying kilns, the effect of non-uniform
airflows
is particularly diff icul t to resolve. Nijdam and
Keey 2002)
have
investigated airflow
patterns
in conven
tional
heat-and-vent
timber kilns to determine design modi
fications that promote more uniform flows. Their velocity
measurements
down the height of the timber stack in a
k i l n wi th outward-swing overhead baffles showed that the
uppermost packets
of the
stack were starved
of airflow.
Nijdam and Keey 2002) demonstrated that, using con
toured right-angled
bends
and inward-swinging
overhead
baffles in higher-velocity wood drying kilns, there was a
3-fold
reduction in the
range
of the vertical velocity
distribution.
In a typical industrial dehumidifier
k i l n ,
the dehumidifier
comprises separate
modules, wi th their own fans, indepen
dent of other fans which may be used to maintain air move
ment in the
k i l n .
In this system the k i l n airflow and the
dehumidifier airflow interact closely, although they are nor
mally not designed as an integrated system. The impact of
mismatches
is diff icul t to anticipate. It is
also
diff icul t to
Correspondence to: Dr Z.F. Sun, Physics Department, University of
Otago, PO Box 56, Dunedin, New Zealand.
E-mail:
measure the distributions of air velocity and
pressure,
due
to the complex interactions of the
subsystems, such
as the
k i l n fans, wood stack, air ducts and the dehumidifier fans
and heat
exchangers.
In this
paper
we
present
an
analysis
of airflow in a commercial dehumidifier wood drying k i l n
as a whole, using a three-dimensional CFD model. We
show
how airflow
design
has the potential to affect the
per
formance of dehumidifier wood drying kilns in unexpected
ways. A modified k i l n configuration, in which an air duct
connects
the exit of the dehumidifier fans wi th the upper
duct space of the
k i l n ,
has
been assessed
using the CFD
model. The
results show
that this ki ln configuration can
significantly
improve the dehumidifier wood drying ki ln
performance. In particular, for dehumidifier wood drying
kilns to achieve high efficiency, it is
advantageous
to use
a single set of
fans
to drive the air flow and to
ensure
all
the flow
passes
the dehumidifiers without recirculation.
SYSTEM DESCRIPTION
Figure 1
shows
the
flow
configuration and the coordinate
system based on a commercial wood drying
k i l n .
The origin
of the coordinate
system
is at the mid-point of the left-hand
w a l l Figure la) of the k i l n at floor level. The system con
sists of two dehumidifier modules installed side-by-side, a
wood stack and six k i l n air recirculation fans, all located
wi th i n an insulated k i l n
chamber.
The two main
elements
of the dehumidifier, which
influ
ence
the flow
patterns
in the
k i l n ,
are the
condenser
and
evaporator
and their air
fans,
shown in Figure 1 a).
Each
module has two
condenser
fans and three evaporator fans.
For simplicity, in this investigation only the characteristics
of the volume flow rate and pressure drops of the
condenser
1344
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A IR
FLOW PATTERNS
1345
5.3 m
4.24 m
kiln fans.
r .
2.2 m-
stack
3.B7m
0.8 m
6.8 m
(a)
U
1.9m
A
air duct
4.3 m
condenser
coils
and fans
evaporator coils
and
fans
0.54 m 1 0 8 m 1.08 m
0.63
m
4.51 m
baffle-board
central 2.75 m 0.5 m
symmetry line
A - A
(b)
B
(c)
Figure
1.
Schematic
diagram of an industrial dehumidifier wood-drying
k i ln .
coils
and fans are considered, since the airflow of the evap
orator enters the
condenser
and the flow rate is typically
only 10% of the
condenser
air f low. Each dehumidifier
module also has an electric air heater to preheat the
k i ln
before drying commences. Here the flow
resistance
of the
heater,
which is located between the condenser fans and
condenser
coils, has
been
incorporated
wi th
the
condenser
coils.
As
shown in Figure 1(a), the cross section of the
k i ln
chamber along the x-direction has a trapezoidal shape,
wi th
the front higher than the rear by 1 m. The backward-
sloping roof would produce stronger vortices in the front
top corner and thus a larger
pressure
drop than the peak-
top and barrel-top kilns discussed by Nijdam and Keey
(2002).
In Figure 1(a), the dashed lines represent the
walls
of a simple air duct which connects the top ceiling
space wi th the exit of the dehumidifier in a modified con
figuration
of
the k i ln .
Stacks of wood are normally several packets deep in the
air
flow
direction and several packets high. As shown in
Figure 1(a),
three
in-line
packets,
being 2.1 m
deep
each
in the air
flow
direction, are considered. The three packets
are supported by bearers 100 mm thick. The gaps
separated
the packets in the airflow direction are 0.25 m wide and
the gaps
separated
the packets vertically are neglected.
Concrete
slabs,
150 mm thick, are placed on the top of
the timber stack to reduce warp during drying (Nijdam
and Keey, 2002). It is
assumed
that the k i l n is fully
loaded
wi th
60 layers of wood boards and thus the horizon
tal ceiling of the drying zone is just on the top of the
concrete slabs.
The wood stack is assumed to be a normally aligned
stack of wood
boards
(Sun and Carrington, 1999), wi th
the geometry shown in detail in Figure 1. It is
assumed
that the side baffles
which
prevent air bypassing the side
of
the timber stack, as shown in Figure 1(b), are fully effec
tive, and air bypass is neglected. The details of the horizon
tal
board layers are shown in Figure 1(c). The
space
between board layers is 20 mm, which is maintained by
longitudinal stickers (or fillets across the width of the
stack at intervals of 450-600 mm (Keey et ai, 2000).
Since the
aspect
ratio of the ducts formed by the board
layers and stickers is relatively large, more than 20, the
effects of the stickers inside the stack on flow patterns
have been neglected. Stickers at the side end of the stack,
however, prevent air flow from the stack to the side
bypass
space.
As demonstrated by previous authors,
the effect of small gaps between in-line boards on the press
ure drop of
airflow
across wood stacks (Langrish and Keey,
1996) and on external mass and
heat
transfer rates (Sun,
2001) can be neglected. Thus, small gaps between in-line
boards have been ignored in the simulation. The ratio of
the inlet plenum-space w id th to the sum of the heights of
the sticker spaces is equal to 0.67 (=0.8/1 .22 m), which
is smaller than the minimum value
(unity)
which was
suggested
by
Nijdam
and Keey (2002) to mitigate the
adverse
effect of the frictional
pressure
drop down the
height of the plenum chamber.
NUMERICAL METHOD
In
order to characterize the distributions of air velocity
and pressure and air recirculation occurring in the k i ln ,
the renormalization group (RNG)
k-e
turbulence model
has been used to solve the turbulent momentum transport
equations. The RNG
k-e
model employs a differential
form of the relation for the effective viscosity,
yielding
an accurate description of how the effective turbulent trans
port varies wi th the effective Reynolds number. This allows
accurate extension of the model to near-wall flows and low-
Reynolds-number or transitional flows (Fluent Incorpor
ated, 1997). The RNG k-e model can be expressed by
the fo l lowing momentum equations and the equations for
turbulent kinetic energy
k)
and its
rate
of dissipation (e)
(Fluent Incorporated, 1997):
9
, , 3 3
(P i) + X - pUiUj)
OX;
/
Peir
3 duj BP
9bc
(1)
d,
„ 9 , ,% 9 /
dk\
p k )
p u t ) = -
U M e f f
- J
+
n , S - p e
and
(2)
3
, N 3
^-(pe) + —(pw,e)
at ax,
3
/ 9e\ , e2
eMeffT- +CUTH,S -C
2
eP-T - R (3)
3.v,
dxi
In
the simulation, the standard RNG k-e model constants,
derived analytically by RNG theory,
have been used:
CM =
0.0845, C
l 6
= 1.42, C
2 e
= 1.68, a
0
= 1.0. The standard
Trans
I C h emE
Part A,
Chemical Engineering Research and Design, 2004,
82(A10): 1344-1352
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A IR FLOW
PATTERNS
1347
300
250
200
0_
150-
£ 100
Q_
50
0-
Woods Air
Movement 2101
Fantech
0714/10
Extension of
fitted relation
2
4 6 8
Volume
flow
rate m3 s~ 1
igure
3
Fan curves of the Woods air movement k i l n fans and the Fantech
0714/10/25 condenser fans.
fo r
the
pressure
rise across each of the
k i l n
fans, and
= 203.029 + 28.7837v - 2.6677v2, 9.1 < v < 15.7
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1348 SU N et al.
Figure 5 shows the velocity field in a vertical y-z plane
coinciding
wi th the
axes
of the condenser fans of the dehu-
midifiers. In this figure, the two horizontally oriented grid-
surfaces represent the condenser fans. It can be
seen
that air
flow eddies are produced at the top side-corners of the k i ln
and
in the downstream
spaces
between the condenser fans.
More
generally, the CFD results indicate that air flow
eddies are produced at all the top and bottom corners of
the
k i ln
and at the corners of the timber packets. The
eddies represent dissipative processes that reduce the effi
ciency
of
both
the
k i l n
fans and the
dehumidifier
fans.
The calculated mass flow
rates
of the two condenser fans
are the same, at 4.89 kg s _ .
Figures 6 and 7 show profiles of the
x-velocity
along
lines from
the stack inlet to stack outlet on two typical
hori
zontal
planes, at the top ( z = 3.71m) and bottom
(z = 0.11 m) of the stack.
I t is
seen from
Figure 6 that the
velocity profiles
on the
top plane z 3.71 m) are not uniform at the entrance
region
of the
first
packet. The x-velocity
spans from
0.5
to 1.5 m s~ approximately. This
variation
in the
velocity
appears to be caused by non-uniform flow in the ceiling
space
of the
k i l n illustrated
in Figure 4. The large range
in the x-velocity indicates that there are large recircula
t ion
eddies in the top entrance
region
of the
first
packet.
Th e flow patterns of the
airflow
on the top plane of the
second packet are similar to those of the first packet, but
the velocities are larger. At the entrance of the
third
packet, the x-velocity span is smaller than those at the
. s . m m
- - > i
i \
• A)
•
• 4 1
1
Figure 5 Vector of
velocity
magnitude on the vertical plane where
x = 8.725 m.
3.50e+00 -|
3.00e+00 -
2.50e+00 -
2.00e+00-
E,
1.50e+00-
?
o
o
1.00e+00-
>
5.00e-01 -
0.00e+00-
-5.00e-01 -
-1.00e+00-
0 1 2 3 4 5 6 7 8
x-coordinate (m)
Figure 6 .r-velocity profiles on the plane z = 3.71 m.
entrances of the first and second packets. However, the
large span in
x-velocity
in the
exit
region
of the
third
packet indicates the presence of recirculation eddies at
the exit of the stack. These appear to be caused by the
k i ln
fans installed in the ceiling space of the
k i l n
and by
the strong resistance of the airflow impelled by the conden
ser fans
illustrated
in Figure 5. Since the effects of the
longitudinal
stickers inside the stack have been neglected
in the simulation, the air velocity spans in the cross y-
direction
may be overestimated. However, this would not
produce serious errors in the average mass flow rates.
The
x-velocities averaged over the cross-sectional
area
of
the top airflow channel at the locations of 1 m from the
leading edge
of each of the packets are 0.94, 1.93, and
2.39 m s~ in the
first,
second and
third
packets,
respectively.
A s shown in Figure 7, unlike the air
velocity
profiles on
the top plane, which increase
from
the
first
packet to the
third packet, the air
velocity
on the bottom plane decreases.
The
average values of the x-velocity in the
bottom airflow
channel are 1.92, 1.59 and 1.37 ms~ in the
first,
second
and third packets, respectively.
2.50e+00
2.25e+00
2.00e+00
„
1.756+00
| 1.50e+00
£ • 1.258+00
o
g 1.00e+00
7.50e-01
5.00e-01
2.50e-01
0.00e+00
. y
= 0
4
- y
= 0.45
• y = 0.9
o
y=1.35
•
o y = 1.8
1
= y = 2.25
a y = 2.7
1 2
3
4 5 6
x-coordinate (m)
Figure 7
.v-velocity
profiles
on the plane z = 0.11 m.
Trans I C h emE
Part A,
Chemical Engineering Research and Design,
2004, 82(A10): 1344-1352
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A IR FLOW
PATTERNS
1349
Th e velocity
profiles
for the
middle
plane of the stack
z
= 1.91 m), which are not shown here, are similar to
the
velocity
profiles at z = 3.71 m and z = 0.11 m shown
in Figures 6 and 7. However, the profiles of x-velocity on
the middle plane of the stack are more uniform than those
on the top plane (z = 3.71 m) and on the
bottom
plane
z
= 0.11 m). The average values of the x-velocity in the
middle
airflow channel are 1.96, 1.88 and 1.89 m s 1 in
the first, second and th i rd packets respectively.
Profiles of the x- and y-velocity along lines from the
stack inlet to stack outlet indicate that, in the
vertical
z-
direction, the velocity
profiles
are also not
uniform.
Figure 8 shows the vertical profiles of the x-velocities aver
aged over the cross-sectional areas of the corresponding
flow
channels at the locations of 1 m from the leading
edge of each of the three packets (solid lines, test 1). It is
seen that, in the first packet,
w i t h
increasing in the height
o f
the stack, the air velocity increases from 1.92 m
s>~
at
the bottom of the stack to 2.20 m s
_
at z = 0.53 m, and
then
decreases to 0.94 m s 1 at the top of the stack. The
vertical
profile of the air
velocity
in the second packet is
more
uniform than that in the first stack, and the velocity
in
the top region is much larger than that in the first
packet. The
vertical
profile of the air
velocity
in the
third
packet shows that the air velocity in this packet gradually
increases w i t h the height of the stack
from
1.37 m s~ at
the
bottom
of the packet to 2.39 njs. at the top of the
packet. The increase in the air velocity in the th i rd packet
may be due to the effect of the k i l n fans, since the upper
part of the th i rd packet is close to the k i l n fans. It appears
that the empty spaces between successive packets have the
effect of
redistributing
the airflow rate through the packets.
The
variations and fluctuations of the
vertical
profiles of the
ai r velocity in the first and second packets are consistent
wi th the measurements by Nijdam and Keey (2002) in a
traditional
peak-top
k i l n
and in a newer barrel-top
k i ln .
Nijdam
and Keey (2002)
found
that the upper part of the
stacks were normally starved of
airflow
at the air entry
end, due to a
recirculation
zone adjacent to the stack
wi th i n the inlet plenum chamber.
A i r
flow recirculation
between the outlet of the dehumi
difier
and its inlet is also illustrated in Figure 9, which
shows pathlines of
massless
particles discharged
from
the
condenser fans. Here, the three vertically oriented gr id-
surfaces represent the
k i l n
fans, the two horizontally
1 2 3
Height
from
base of timber stack (z-direction) (m)
igur 8 Ai r velocity profile along height of the stack.
igur 9 Illustrative pathlines of particles leaving condenser fans. Par
ticles are coded by the grey colour.
oriented grid-surfaces denote the condenser fans, and the
grid-box represents
the
dehumidifier
walls.
This
figure
shows that a significant fraction of air leaving the dehumi
difier
mixes w i t h the
airflow
out of the stack and re-enters
the
dehumidifier,
because the k i l n fans and the condenser
fans are not configured in an integrated way. Indeed, part
o f the airflow recycles several times from the outlet of
the dehumidifier to its inlet.
Overall calculated results for the k i l n
w i t h
the present
configuration (Figure 1) are listed in Table 1 along w i t h
results for a modified configuration discussed below. In
Table
1, ws is the x-velocity averaged over the whole
cross-sectional area of the stack at the locations inside
each of the three packets, 1 m from the inlet of each
packet. It is seen that, for this
k i ln
(test 1 in Table 1), the
total mass
flow rate (19.55 kg s~ ) delivered by the conden
ser fans is larger than that (15.70 kg s_ ) delivered by the
k i l n
fans. The difference (3.85 kg s~ ) between the
mass
flow
rates delivered by the condenser fans and k i l n fans
represents the minimum amount of
airflow
recirculation
from
the
exit
of the dehumidifier to its
inlet.
This would
happen i f all the
airflow
leaving the stack were to enter
the dehumidifier. For the present
k i l n ,
the minimum
amount
of the airflow
recirculation
is 19.7% of the total
ai r mass
flow rate delivered by the condenser fans. How
ever, the
simulation
indicates that the real
recirculation
mass flow rate is l ike ly to be much larger than this
m i n i
mum
value, since much of the
airflow
delivered by the
k i l n fans comes from the stack directly.
Figures 10 and
11
show the pathlines
o f massless
particles
which are discharged from the upper and
lower
rear-end
surfaces of the stack respectively. In Figures 10 and 11,
the vertical rectangular gray surfaces represent the rear-
end
surfaces of the upper 21 board layers and lower 39
board layers of the stack, respectively. It can be
seen from
Trans I C h emE
Part A, Chemical Engineering Research and Design 2004, 82 A10): 1344-1352
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1350 SUN
et
al.
Table
I Calculated results for different k i l n configurations.
Number of Number dehumidifier
AP
wk fan
Test
k i l n
fans modules Duct
(Pa)
k g s
')
k g s
')
(kgs
- 1
)
( ms
_ 1
)
1
6
2 4
condenser fans) No 31 34
15.70 19.55 15.70 <
1.91
2
0
2 4
condenser fans)
Yes
38.58
18 26
18 26
2.23
Figure
10 that all the air flow discharged
from
the f low
channels of the upper 21 board layers of the stack,
wi th
some local recirculation
in
the space above the dehumidifier,
enters the
k i l n
fans
directly.
Figure 11 shows that, although
most of the air flow discharged from the lower 39 board
layers enters the condenser and then passes
across
the dehu
midifier fans, part of the air flow discharged from the
lower
part of the staGk
goes
to the
k i l n
fans
directly.
Th e mass flow rate, 6.15 kg s~', of the air discharged
from
the upper 21 board layers of the stack is 39.2 of
the
total
flow rate of the
k i l n
fans. Hence
less
than 60.8
o f
the total flow rate of the
k i l n
fans comes from the
dehumidifier. Thus, using the data shown in Table 1 for
test 1, the mass flow rate entering the k i l n fans from the
dehumidifier is
less
than 9.55 kg s~' and the
mass
f low
rate of air
recirculation
from the exit of the dehumidifiers
to
its inlet is more than 10.0 kg s_ 1 , representing 51.2 of
the
total
flow rate of the condenser fans. This
airflow
recir
culation would raise the temperature and reduce the humid
it y
of air at the inlet of the dehumidifiers, which would
reduce the efficiency of the dehumidifiers. In
addition,
energy used by the condenser fans to maintain this large
amount of recirculation f low, approximately 51.2 of the
total energy used by the condenser fans, does not make
any
contribution
to the performance of the
k i l n .
MODIFIED CONFIGUR TION
To
illustrate the effect of changing the dehumidifier kilns,
a modified configuration, test 2 in Table 1. was investigated
using the CFD turbulence model. In test 2, air recirculation
from
the
exit
to the
inlet
of the dehumidifier was e l im i -
nated, using an air duct connecting the dehumidifier to
the top ceiling space of the
k i l n ,
shown by the dashed
lines
in Figure 1(a).
The vertical profiles of the average x-velocity in test 2
are shown by the dashed lines in Figure 8. The air velocities
in the three packets of the k i l n in test 2 are significantly
larger than those in test 1. In addition, the velocity distri-
butions
from
the
first
packet to the
th i rd
packet are more
uniform than those in test 1 and the air velocity profiles
in
the
vertical
z-direction in test 2 are
similar
to those in
test 1. This indicates that, by using appropriate ducting
and
only the condenser fans, the drying performance of
the
k i l n would
be improved.
Overall
calculated results for test 2 are also
listed
in
Table 1. It is seen that, without using
k i l n
fans but wi th
an air duct, the calculated average
x-velocity
(2.23 m s
_
)
in test 2 is larger than that (1.91 m s- 1 ) in test 1. Since
no
k i l n
fans are used, the energy consumption of the
k i l n
fans would be eliminated in test 2. As shown in Table 1,
compared w i th test 1 the
overall mass
flow rate of the
Figure 10
Pathlines of the particles discharged
f rom
the rear-end surface
o f the upper
21
board layers of the stack. Particles are coded by the grey
colour.
Figure
11
Pathlines of the particles discharged
f rom
the rear-end surface
o f the lower
39
board layers of the stack. Particles are coded by the grey
colour.
Trans I C h emE
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A IR FLOW PATTERNS
1351
condenser
fans in
test
2 has decreased by 1.29 kg s~ and
the stack pressure drop in
test
2 has increased by 7.4 Pa.
The volume flow
rate
of
each condenser
fan has
decreased
from 4.07 to 3.80 m3 s~ , and the power of the
condenser
fans
should
increase
by approximately 1 . On the other
hand, the adverse implications of the air recirculation on
the performance of the dehumidifier, which are
discussed
below, do not occur in test 2.
INFLUENCE OF RECIRCUL TION ON DRYER
PERFORM NCE
The CFD turbulence model cannot be
used
to simulate
the thermodynamic cycle and mass and heat transfer pro
cesses in the dehumidifier or in the wood boards. Hence,
to
assess
the influence of
airflow
recirculation at the dehu
midifiers on the performance of the
k i ln
system, a dynamic
dehumidifier wood drying model developed previously
(Sun et al. 2000) has been used. The model solves the inte
gral form
of the
unsteady
state mass, momentum and
energy balance equations
for both the air flow and the wood
boards.
The distributions of the
average mass
fractions,
temperature, pressure and velocity of the air
stream,
as
wel l as the average moisture content of the wood boards
and their temperature, can be estimated using this model.
For illustrative purposes, the timber is Pinus
radiata
sap-
wood
wi th an
i n i t i a l
moisture content 140 . The final
moisture content is 13 , and the volume of timber is
70 m3 . There are two dehumidifier modules in the k i ln ,
and the external heat and
mass
transfer
rates
between air
flow
and the surfaces of wood boards were calculated
using correlations established on the basis of modified
boundary layer
theories
which
take account
of the
separ
ation and reattachment flows (Sun, 2002). The maximum
condensing temperature, maximum evaporating
tempera
ture, and minimum
stack
inlet relative humidity are
l imi ted
to be 75 °C , 25=C and 40 respectively (Carrington et al
1995). An air
preheater
of 30 kW is
used,
which is turned
of f when the stack inlet dry-bulb temperature reaches
50°C . The
stack
inlet air velocity is 2 m s ' , based on the
CF D results as shown in Table 1 for the
present
k i l n (test 1).
The calculated overall dehumidifier
energy
use,
k i ln
fan
energy, heater energy, and drying time are shown in
Figures 12 and 13, respectively. It is
seen from these
figures
1.6
1.5
f 1.4
Z
g
* 1.3
>,
at
CD
c 1 2
1 1
-
8/18/2019 Air Flow Patterns
9/9
1352
SUN et al.
22
141 , , 1
14 16 18 20
Model Power (kW)
Figure
15. Comparison of
measured
dehumidifier power consumption
wi th
a model benchmark under conditions
w i t h
and witho ut recirculation.
the temperature and humidity when the measurements
were made. This difficulty is avoided by comparing the
mea
sured data
w i th
dehumidifier model predictions, the model
assuming no recirculation, to provide a common perform
ance
reference for the
measured data.
Accurate
agreement
wi th
the model is not necessary for this comparison. In the
results, shown in Figures 14 and 15, the
measured
drying
rate is consistently below the model prediction when there
is no baffle, but close to the model results when the baffle
is present. In addition the data indicates that the power
use is higher when the baffle is not present.
CONCLUSION
A i r flow
patterns in an industrial dehumidifier wood
drying k i ln
have been
investigated using a CFD model. In
order to solve the computational
difficulties
for simulation
of
a practical k i ln a simplified procedure has been devel
oped in which the spaces between the board layers are trea
ted as an isotropic porous medium. A momentum source
term describing the pressure drop w i th in the porous
medium
has been added to the standard
fluid flow
equations.
The results obtained show that, without suitable air duct
ing at the dehumidifier air discharge, a large fraction of the
dehumidifier
airflow recycles back to the inlet of the dehu
midifiers. Using a dehumidifier wood drying k i l n model, it
has been demonstrated that such air recirculation
reduces
the efficiency of the dehumidifiers and increases drying
time,
by 18 and 14%, respectively, for the example
presented.
A modified
k i ln
configuration, in which an air duct con
nects the dehumidifier
w i th
the upper
airflow
channel of the
k i l n has been analysed. The results show that this configur
ation
significantly improves the dehumidifier performance.
I t is concluded that, for dehumidifier wood drying kilns to
achieve high efficiency, it is important to (a) ensure the
airflow is properly ducted to prevent recirculation and
(b) avoid using two
sets
of air fans in series.
NOMENCL TURE
k turbulent kine tic energy (m s~~)
M
mass flow
rate (kg s )
P
pressure (N trT
-
)
A f s pressure drop of air
flow
passing through stack (N m~
2
)
S/
momentum source term for porous media (N m ~ )
r time (s)
it veloci ty component or velocity in jr-direction (m s ')
s average x-velocity in stack (m s~ )
v velocity (m s
_ l
)
x coordinate (m)
coordinate (m)
z coordinate (m)
Greek symbols
ak inverse effective Prandtl number for turbulent kinetic
energy
ct€ inverse effective Prandtl number for turbulent dissipa
t ion rate
e turbulent dissipation rate (m
2
s
- 3
)
iu eff effective viscosity p., + p.) (kg m
-
' s
- 1
)
fj. y turbulent viscosity (kg m _ 1 s_ l )
p
density (kg m~
3
)
Subscripts
c,fan condenser fans
i /-direction
j /-direction
k k i l n
k,fan
k i l n
fans
REFERENCES
Carrington, C.G., Bannister, P. and Li u, Q., Performance analysis of a
dehumidifier using HFC-134a, Int J Refrig, 18: 477-485.
Fantech Pty Ltd, 1993.
Fans
by
Fantech,
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Fluent Incorporated, 1998,
GAMBIT
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CHEMECA
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Sun, Z.F. and Carrington, C.G., 1999, Effect of stack configuration on
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drying
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CKNOWLEDGEMENTS
The authors gratefully acknowledge the New Zealand Foundation for
Research Science and Technology for supporting this work under contract
UOOX0004.
The
manuscript was received I July 2002 and accepted for publication
after revision 18 June 2004.
Trans I C h emE
Part A,
Chemical Engineering Research and Design
2004, 82 A10): 1344-1352