Group Title: 7th International Conference on Multiphase Flow - ICMF 2010 Proceedings
Title: 13.6.4 - Inhomogeneous Compressible Numerical Analysis with Phase Change of the Mixing Phenomenon in Two-Phase Ejector using CO2
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Permanent Link: http://ufdc.ufl.edu/UF00102023/00334
 Material Information
Title: 13.6.4 - Inhomogeneous Compressible Numerical Analysis with Phase Change of the Mixing Phenomenon in Two-Phase Ejector using CO2 Multiphase Flows with Heat and Mass Transfer
Series Title: 7th International Conference on Multiphase Flow - ICMF 2010 Proceedings
Physical Description: Conference Papers
Creator: Nakagawa, M.
Marasigan, A.
Matsukawa, T.
Publisher: International Conference on Multiphase Flow (ICMF)
Publication Date: June 4, 2010
 Subjects
Subject: inhomogeneous
compressible
Eulerian-Eulerian model
two-phase ejector
phase change
 Notes
Abstract: The effectiveness of the two-phase ejector is highly dependent on the mixing characteristics of high-speed motive flow and suction flow. It should be further studied to make the use of ejector particularly in CO2 refrigeration cycle more viable. In this study, numerical simulations of the mixing and pressure recovery process inside a two-phase CO2 ejector were carried out at different operating conditions and verified with experimental data. The effects of compressibility and mass transfer due to phase change were included. The flow regime was consisted of dispersed droplets in continuous vapor. The numerical results have shown good agreement with the experimental data. The pressure recovery profile and its magnitude were largely affected by the motive nozzle inlet conditions. The suction inlet pressure yielded minimal effect on these parameters. The effect of phase change was considerably small for the cases used in this research while the energy conversion efficiency of the motive nozzle has increased for increasing inlet pressure and lower inlet temperature. It confirmed our conclusion in our previous study that this efficiency, as well as the coefficient of performance of the system, increases as the inlet condition moves to the left of a P-h diagram. The results of this study provide useful information for ejector optimization and better understanding of the process taking place inside a high-speed two-phase CO2 ejector.
General Note: The International Conference on Multiphase Flow (ICMF) first was held in Tsukuba, Japan in 1991 and the second ICMF took place in Kyoto, Japan in 1995. During this conference, it was decided to establish an International Governing Board which oversees the major aspects of the conference and makes decisions about future conference locations. Due to the great importance of the field, it was furthermore decided to hold the conference every three years successively in Asia including Australia, Europe including Africa, Russia and the Near East and America. Hence, ICMF 1998 was held in Lyon, France, ICMF 2001 in New Orleans, USA, ICMF 2004 in Yokohama, Japan, and ICMF 2007 in Leipzig, Germany. ICMF-2010 is devoted to all aspects of Multiphase Flow. Researchers from all over the world gathered in order to introduce their recent advances in the field and thereby promote the exchange of new ideas, results and techniques. The conference is a key event in Multiphase Flow and supports the advancement of science in this very important field. The major research topics relevant for the conference are as follows: Bio-Fluid Dynamics; Boiling; Bubbly Flows; Cavitation; Colloidal and Suspension Dynamics; Collision, Agglomeration and Breakup; Computational Techniques for Multiphase Flows; Droplet Flows; Environmental and Geophysical Flows; Experimental Methods for Multiphase Flows; Fluidized and Circulating Fluidized Beds; Fluid Structure Interactions; Granular Media; Industrial Applications; Instabilities; Interfacial Flows; Micro and Nano-Scale Multiphase Flows; Microgravity in Two-Phase Flow; Multiphase Flows with Heat and Mass Transfer; Non-Newtonian Multiphase Flows; Particle-Laden Flows; Particle, Bubble and Drop Dynamics; Reactive Multiphase Flows
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Bibliographic ID: UF00102023
Volume ID: VID00334
Source Institution: University of Florida
Holding Location: University of Florida
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Resource Identifier: 1364-Nakagawa-ICMF2010.pdf

Full Text


Paper No 7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30-June 4, 2010


Inhomogeneous Compressible Numerical Analysis with Phase Change of the Mlixing
Phenomenon in Two-Phase Ejector using CO2


Masafumi Nakagawa, Ariel Marasigan* and Takanori Matsukawa


Department of Mechanical and Structural System Engineering, Toyohashi University of Technology,
1-1 Hibarigaoka, Tempakucho, Toyohashi City, Aichi, 4418580, Japan
*Email: ariel iinak: mech tut ac jp


Keywords: inhomogeneous, compressible, Eulerian-Eulerian model, two-phase ejector, phase change

The effectiveness of the two-phase ejector is highly dependent on the mixing characteristics of high-speed motive flow and
suction flow. It should be further studied to make the use of ejector particularly in CO, refrigeration cycle more viable. In this
study, numerical simulations of the mixing and pressure recovery process inside a two-phase CO, ejector were carried out at
different operating conditions and verified with experimental data. The effects of compressibility and mass transfer due to
phase change were included. The flow regime was consisted of dispersed droplets in continuous vapor. The numerical results
have shown good agreement with the experimental data. The pressure recovery profile and its magnitude were largely affected
by the motive nozzle inlet conditions. The suction inlet pressure yielded minimal effect on these parameters. The effect of
phase change was considerably small for the cases used in this research while the energy conversion efficiency of the motive
nozzle has increased for increasing inlet pressure and lower inlet temperature. It confirmed our conclusion in our previous
study that this efficiency, as well as the coefficient of performance of the system, increases as the inlet condition moves to the
left of a P-h diagram. The results of this study provide useful information for ejector optimization and better understanding of
the process taking place inside a high-speed two-phase CO, ejector.


(Sriveerakul et al. 2007, Varga et al. 2009), air (Yaday and
Patwardhan 2008, Hemidi et al. 2009), and Freon-based
refrigerants (Rusly et al. 2009, Zhu et al. 2009, Bartosiewicz
et al. 2006). To our best knowledge, there has been very
limited published research on the CFD modeling of
two-phase ejector using CO2. We have been researching on
CO? ejector through experimentation and numerical
modeling (Nakagawa et al. 2004, Nakagawa et al. 2009a,
Nakagawa et al. 2009b). One of our research objectives is to
model numerically the dynamics inside the ejector using
CO2 as the working fluid.
The geometrical features of an ejector, shown in Figure
1, for CO2 refrigeration application are similar to the ejector
using other working fluids. However, the inlet of the motive
nozzle for the case of a CO? ejector is at transcritical state,
as shown in Figure 2. The fluid then expands to two-phase
state inside the motive nozzle. This process generates
high-speed dispersed droplets in continuous vapor flow. The
outlet flow of the motive nozzle then entrains and mixes
with the suction vapor flow from the evaporator to facilitate
pressure recovery from the mixing section to the diffuser. A
full CFD model of this whole process is desirable but it
involves complex processes such as the expansion from the
transcritical state which is an obstacle to attain that objective.
This study focuses on the mixing phenomenon inside the
ejector using CO2 including compressibility effects and
phase change. The numerical results were verified with the
experimental data. The effectiveness of atwo-phase ejector
is highly dependent on the mixing characteristics of
high-speed motive flow and suction flow. A numerical
model of this complex process is essential in designing an
optimized ejector.


Introduction

The simple design of ejectors and its potential in
utilizing available energy make it practical for wide range of
applications such as power plants and aircraft. In a standard
vapor refrigeration cycle, the available energy dissipated in
the throttling process of the working fluid can be utilized by
replacing the expansion valve with a two-phase ejector. Its
application in this field extends for different working fluids
but due to the worsening problem of global warming, it is
desirable to focus the study of two-phase ejector using
alternative refrigerants. The natural refrigerant CO?. which
has zero ozone depletion potential and very low global
warming potential, is seriously being considered today to
replace conventional refrigerants. However, significant
amount of energy is lost during expansion process when it is
used in conventional refrigeration system. The use of
two-phase ejector can address this problem and it should be
studied further to make the use of alternative natural
refrigerant CO2 practical and efficient. Experimentation and
numerical modeling aid in achieving this goal. The latter is
desirable to give a better picture of the dynamics of the
ejector operation. However, any numerical modeling should
be verified with experimental data.
One of the classical 1-dimensional models of ejectors
was formulated by Keenan and Neumann (1942) using air as
the working fluid. Since then, a number of researches based
on thermodynamic and dynamic numerical modeling of
ejectors from 1-dimensional to 3-dimensional approaches
were published (He et al. 2009) In the advancement of
computational machines and codes, the use of computational
fluid dynamics (CFD) technique is deemed as more
appropriate to capture the actual flow inside the ejector.
Most of the researches based on CFD have used steam





























1 Gas cooler
L '


5 o-


Figure 3: Ejector assembly used in the experiment




Table 1: Locations of the thermocouple along the along the
longitudinal axis of the ejector with x=0 at mixing outlet

No. x (mm) No. 15mm

1, 2 suction nozzle 11 3 (mix)
3 -21.35 (conv) 12 6.7 (mix)
4 -18.75 (conv) 13 10.3 (mix)
5 -15.55 (conv) 14 14.6 (mix out)
6 -9.9 (div) 15 19.1 (dif)
7 -8.05 (div) 16 25.6 (dif)


7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30-June 4, 2010

EXPERIMENTAL METHOD

In this study, an ejector with rectangular cross-section
WaS employed. This research was carried out for academic
7 purposes and this ejector design is easier to fabricate as
compared to circular type. The exploded view of the
ejector assembly is shown in Figure 3. The bottom part
WaS a 20mm thick bronze plate where holes were drilled as
Suction and motive nozzles inlets and an ejector outlet. 19
K-type thermocouples were placed on the top wall along
the longitudinal axis of the ejector. These were fixed using
epoxy inside the 00).8mm holes drilled on the top Bakelite.
During experimentation, these thermocouples measured the
static temperature of the fluid trapped inside the 00).8mm
holes. Theoretically, the temperature readings, specifically
from mixing section to diffuser, can be used to obtain the
pressure profile in these sections since the flow is in
two-phase state. In order to verify this, pressure gauges
were fixed instead of thermocouples and the system was
run at similar operating conditions. The results showed that
the saturation pressure based on thermocouple readings
were comparable to the pressure transducer measurements.
The horizontal positions of the thermocouples used in this
study are shown in Table 1 with origin set at the inlet of
the mixing section.


Paper No


MOOE





FLOW


DIOUENR


Figure 1: Illustration of a two-phase ejector


- Ejector Cycle
----- cntorwntnl Ccol*


CO,


Figure 2: P-h diagram of an ejector refrigeration
cycle using CO2


p. steel

SBakelite






SBake le

Bronze

x J U

z:


Epoxy


Symbol
th
g
h
IAD



P

T
TPC
U



Subscript
1
2
3
4


Quantity
nozzle mass flow rate
entrainment ratio
enthalpy
inlet section from motive
nozzle
inlet section from suction
nozzle
pressure
slip ratio
temperature
Thermal Phase Change
axial velocity
quality
density
efficiency

motive nozzle inlet
motive nozzle outlet
suction nozzle inlet
suction nozzle outlet
mixing inlet
diffuser inlet
diffuser outlet
gas cooler outlet
converging nozzle
diffuser
diverging nozzle
motive nozzle
maximum
evaporator
ejector
profile
recovery
suction nozzle


SI Unit
kg/s

kJ/kg


EJE top plate

EJE center plate

EJE bottom plate 4


SUCTION IN
MOTIVE IN
SUCTION IN


kg/m3


S10
2 8



!4


Enthalpy (kJ/kg)






7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30-June 4, 2010

pressure and temperature.
WATER IN WATER OUT


Paper No


The main ejector assembly was composed of three
stainless plates and the schematics of the center plate are
shown in Figure 4. The motive nozzle had a throat width of
0.4mm, divergence angle of 0.970 and divergence length of
11.5mm. This nozzle design was based on our previous
experimental data and numerical analysis using isentropic
homogenous equilibrium model (Nakagawa and Morimune
2003). The ejector mixing area ratios which is defined as
the ratio of mixing area to the motive nozzle outlet area
was 4.22. The mixing length was 15mm while the diffuser
section had divergence angle and length of 5.20 and
30.5mm, respectively. The basis of this design was also
from the result of our previous study using 1-dimensional
3-velocity numerical model of the mixing process for R-12
ejector system (Nakagawa et al. 1994). The whole ejector
assembly was tested for leakage at a pressure of at least
10MPa before installing it in the main system.




O O O O


R3 5

15 O



re In millimeter


O
x


( O

dimensions a


Figure 4: Schematics of the center ejector metal plate


The experimental setup used in this study, as shown in
Figure 5, was a modified conventional refrigeration system
with heat exchanger. An internal heat exchanger with an
effectiveness of about 35% was employed between the
inlet of the compressor and outlet of the gas cooler.
Pressure transducers and thermocouples were placed
between system components to monitor and maintain the
desired operating conditions and to obtain the required
parameters to evaluate the fluid properties. The suction
flow rate sh,~ was obtained based on the enthalpy
difference of the air passing through the evaporator. The
required parameters for the air flow were measured using
thermocouples, differential pressure transmitter and
hygrometer. The motive flow rate shz was calculated
using heat balance in the gas cooler where circulating
water was used as heat sink. The evaporator fan and
compressor were controlled using inverters. The accuracies
of these measuring devices are summarized in Table 2. The
experiments were carried out at different operating


AIR IN I~T
S: pressure transducer T* : thermocouple

Figure 5: Schematic diagram of the experimental setup


Table 2. Accuracy of measuring devices used in the
experiment
Measuring Device Accuracy
Pressure transducer +().25% full scale
Thermocouple +().5oC actual reading
Differential pressure transmitter 1.()% full scale
Hygronieter 2% actual reading



NUMERICAL MODEL

Domain Setup

The numerical calculation was carried out using the
commercial software ANSYS CFX11. The details of the
numerical models that will be presented can be found from
the software's documentation (ANSYS User's Guide v.11).
The inhomogeneous steady-state Eulerian-Eulerian
approach was used and the flow regime was set as dispersed
liquid droplets in continuous vapor. The Particle model was
used for modeling interphase transfer. The droplet was
assumed to have a constant diameter of 10pLm which was
estimated from the data of Padorski et al. (1980). The
k-epsilon and Schiller Naumann models were employed to
model the turbulence effects and interphase momentum
transfer, respectively. The Total Energy heat transfer model
was employed to include compressibility effects. The CO,
fluid properties were obtained using Redlich Kwong
equation state provided in CFX. The simulations were
carried out for both cases of with and without mass transfer
'vhich was modeled using Thermal Phase Change Model
and the Two Resistance Model was used to calculate the
heat transfer coefficient.






7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30-June 4, 2010


Paper No


INS1


IND m



IN S2


Figure 6: Computational mesh


The computational mesh, as shown in Figure 6, was
generated using the ANSYS ICEM CFD code. The control
volume of the CFD model was primarily consisted of
mixing section and diffuser to focus on the dynamics of the
mixing phenomenon of the motive and suction flows. This
domain is represented by the region enclosed by broken
lines in Figure 4. The mesh was constructed such that the
orthogonality, expansion factor and aspect ratio were within
the acceptable range. Several geometrical meshes with
increasing number of elements were first used in this study
until grid independence was obtained. Symmetric
assumption, which is not obvious because of the geometric
design, was not employed in this study to better capture the
actual physics of the flow.



Boundary Conditions

The inlet section IND of the fluid flowing from the
motive nozzle was located 2mm before the mixing inlet or at
the location of the last thermocouple along the motive
nozzle's longitudinal axis. The inlet condition for this
section was set to mixed subsonic and supersonic flow. In
this case, the temperature and pressure must be set which
were obtained from the experimental data. Inlet velocity of
the fluid was also set at this section. The mean velocity of
the each phase was calculated based on the following
equation:


U, (1)
(s2 1)x +1

Where, the sub scripts 1 and 2s correspond to the actual inlet
and isentropic outlet of the motive nozzle, respectively. The
parameters h, s, iln, and x denote enthalpy, slip ratio, motive
nozzle isentropic efficiency and quality. A constant slip ratio
equal to 1 was assumed at this inlet section. Several values
of motive nozzle efficiency were used in the calculation to
obtain the pressure profile and compare it with experimental
data. The vapor void fraction was calculated using the
following relation:


a = (2)
1+1- xX JP s ;7



In order to better represent the fluid velocity at IND, a
non-uniform velocity was employed using the 1/7" power


law which
expressions:


was calculated


based on the following


U (y. z) = Umax


-, u = ( -i+


Where t and b are the thickness and base, respectively, of
section ID and n was set to 7.
Meanwhile, the location of the inlet section INS which
corresponds to the inlet of the fluid flowing to the domain
from the suction nozzle was decided based on some initial
calculations. At first, the whole actual suction nozzle was
modeled and the results were compared with that of using
the grid presented in Figure 6. The results showed negligible
difference so that the latter was used for other calculations.
The inlet condition for this section was set to mass flow rate
using the experimental dataliz~and the flow was assumed to
be pure vapor. The inlet temperature at this section was also
set based on experimental data.
The outlet boundary condition was defined using the
pressure obtained from experiment. The wall was set to
adiabatic and the wall influence on the continuous and
dispersed phases was set to no slip and free slip, respectively.
The convergence was judged by lowering the residuals until
le-5, monitoring pressure and temperature at different
points until it stabilizes, and until the mass imbalance
reaches the value of le-5 kg/s. However, recirculation zone
near the outlet of the diffuser section was obtained at high
inlet pressure P,,, which made convergence based on
residuals difficult to achieve. In such case, the stability of
the values of monitored points and imbalance can be used as
convergence criteria (ANSYS USER's Guide).



RESULTS AND DISCUSSION

The CFX results were validated by comparing the
magnitude of suction inlet pressure and the pressure
recovery profile inside the ejector with the experimental
data. In a trancritical CO, ejector refrigeration cycle, the
three operating parameters; motive nozzle inlet pressure or
the gas cooler pressure P,,,, gas cooler outlet temperature T,
and evaporator temperature Te are important in accessing the
performance of the system. The numerical simulation for
each set of operating conditions was carried out for the cases
Of without mass transfer (noTPC) and with mass transfer


L






























IIIIIIIIIII


0 10 20 30 40 50
x [mm]


7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30-June 4, 2010


Paper No


using Thermal Phase Change Model (TPC). In addition to
that, each case was performed using different motive nozzle
inlet velocities which were calculated based on the assumed
values of motive isentropic efficiency im. The following
graphs show the comparison of experimental suction
pressure and experimental pressure recovery profile with the
calculated results. These values for numerical results were
connected from suction inlet to mixing section inlet for
better visualization. The other lines indicate the centerline
velocities of gas and droplet phases as a function of the axial
distance and only the case of with mass transfer model was
included in this research.


-u -7 U.1.1 TPC Papmm 11 5%.nTPC
u =1,70%, TPC -.Ppme ilm=75% nTPc
- -- -~~ u=70% TPC m-%nTc
O3



I Pn=9MPa
~~To=45oC Te=5"C

--
------


300

250

200 C

150 &i

100

so


4.4



a 4.2


4.0


30 40 50


0 10 20
x [mm]
(a)


Varying Motive N~ozzle Inlet Pressure P,,,


_ -u~~~% TPC
u -.=80% TPC
- -U~ 1=80%. TPC


* -Papenm 1=5%, noTPc
-- P r~ '1m=85%. TPC
--- -P~, ror =80%, noTPC


The results for varying inlet pressure Pn from 9MPa to
11MPa are shown in Figures 7a to 7d. In this case, the gas
cooler outlet temperature To and evaporator temperature Te
were kept constant at 450C and 50C, respectively. The
numerical results show good agreement with respect to the
experimental pressure data. The difference between the
results of with and without mass transfer was almost
negligible for both pressure and velocity profiles. It suggests
the minimal effect of mass transfer in the mixing process
inside a two-phase CO2 ejector.
The motive nozzle efficiency rlm was varied from 70%
to 90% for the range of Pn used in this section. This was
done since the motive nozzle outlet velocities of gas and
droplet phases are not well established. The basis of the
assumed values of rlm was the experimental suction pressure
data and the experimental pressure recovery profile. It can
be seen from the graphs that the total pressure recovery Pre,
which is the difference between the suction inlet and ejector
outlet, changes with respect to the value of im. The total
pressure recovery increased as rlm increased because of
higher energy input at higher values of im. The value of rm
which would best fit the experimental data curve has
increased for increasing value of Pn. It implies the higher
efficiency of motive nozzle in converting pressure energy to
kinetic energy at higher inlet pressure.
In the values of Pn used for this section, the numerical
simulation for 10MPa and 10.5MPa yielded excellent results
in comparison with the experimental data. At lower
P,,=9MPa, the numerical results followed the experiment
data except the region near the mixing outlet and diffuser
inlet. The pressure profile from numerical calculation was
higher than the experimental data which means that the
kinetic energy at this region was converted to pressure
energy more efficiently than the actual process. The trend
was different for higher P,, as shown in Figure 7d. In this
case, the numerical simulation yielded better results except
the inside region of the mixing section. It means the
conversion from kinetic to pressure energy was insufficient
using the numerical simulation.
The results for centerline velocity profile for the cases
of with and without mass transfer were almost the same so
that only the numerical results with mass transfer model are
shown in the graphs. In all of the values of P,, used in this
section, the centerline velocities of liquid and gas phases
had negligible difference. Small velocity difference was
obtained near the inlet of the mixing section but this has
disappeared for higher Pan-


250


200

c
150 3

100


50


-- -P~~ 1m=80%, TPc




Ic IXn= 10MPa
u, T = 450C Te= 50C


0- 4.2

an


3.8 E


.. -u ~l=9 %, TPC
- U n =90%. TPC
- -Up a =80%. TPC
- u om=80%. TPC


e penment
fror~, r~m=904
- P or, 11 =904
- P~ orl, a=809


250


200

C
150


100


So


4q.2


0-4.0


1
30 40 50


0 10 20
x [mm]
(c)


- -U. 11=90%. TPC
- U =90%. TPC
---u.p=80%,TPC
-- 4 n,,=80%, TPC


* P

-- P~~ ,,=904
--P ~~=0


250


200


150 .-


100


so


,2

,4.0


30 40 500


0 10 20
x [mm]


Figure 7: Comparison of numerical results and experimental
data and the calculated horizontal centerline gas velocity U,














































- -U. 11m=90%, TPC experiment
- u~, =90%. TPC --- P,.,ir m=90%, noTPc -
- -U up =80%, TPC -Ppror m,=90%, TPc
- u~,m=80% TPC -'--,i -~8)%. noTPC -
.- ~1,=0. C

'z -r I

j~ jP p~= 10MPa
W. To=420C T= 20C

- -


0 10 20 30 40 5
x [mm]
(a)


7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30-June 4, 2010


Paper No


and liquid velocity UI for varying motive nozzle inlet
pressure Pn at (a) 9MPa, (b) 10MPa, (c) 10.5MPa and (d)
11MPa


- ul,, =90%. TPC
- -U Em=90%,. TPC
- u.1=80%. TPC
- -U~ Im=80%,TPC


- P,

--Pn


xpenment
ror '1m=90%. noTPC
ror 1im=90%. TPC
ror 1m=80%, noTPC
ror Em=80%. TPC




P= 10MPa
= 420C T="


250


200


150


100


so


a


24.4


4.2


4.0


Varying Evaporator Temperature Te

The results for varying evaporator temperature Te from
20C to 80C at P,=10MPa and T,=420C are shown in Figures
7a to 7d. The numerical results again show good agreement
with respect to the experimental pressure data. The
negligible effect of including the mass transfer in the
calculation was also obtained. This trend was obtained for
both pressure and velocity profiles.
The numerical simulation for this section was only
carried out for values of rlm equal to 80% and 90%. This was
enough to obtamn results comparable to the experimental
data. It shows that the suction pressure has minimal effect
on the motive nozzle performance which is logical since the
inlet conditions of the motive nozzle were not varied.
However, it is interesting to point out that at these different
values of suction pressure, the entrainment ratio had
negligible difference as obtained from the experiment.


c eh


ST.


30 40 50


I_


0 10 20
x [mm]


Figure 8. Comparison of numerical results and experimental
data and the calculated horizontal centerline gas velocity U,
and liquid velocity UI for varying evaporator temperature Te
at (a) 20C, (b) 50C and (c) 80C

The difference between the experimental and numerical
results of pressure profile can be seen near the central region
of the mixing section for all values of Te. It should be noted
that although the inlet pressure is at 10MPa, the temperature
To is lower as compared to the case of varying pressure Pn.
The profiles of both experimental and numerical had almost
maintained a constant shape even though the suction
preSsure has changed. The only difference is the magnitude
of the minimum and maximum pressure for each graph. It is
in contrast for the case of varying Pn where the shape of
pressure recovery had changed for a different value of Pn.
Also, the difference between the centerline velocities for
each phase was negligible.



Varying Gas Cooler Outlet Temperature To

The actual inlet temperature of the ejector was lower
than the gas cooler outlet temperature Te since internal heat
exchanger was employed between these two points. The
average temperature drop at the high pressure side of the
heat exchanger was about 7 degrees. The effect of
temperature To can be obtained by analyzing Figure 7b and
8b. In these two graphs, the pressure Pn and temperature Te
were kept at 10MPa and 50C, respectively while the
temperature To was changed from 420C to 450C.
The value of rlm which gave better numerical results in
comparison with the experimental total pressure recovery
was higher than for the case of lower Tc. However, the
numerical results for higher To yielded better approximation
of the pressure curve. At lower T,, the numerical results of
the pressure recovery profile inside the mixing section had
sginificant deviation with the experimental data. The slip
velocity near the inlet of mixing section was negligible for
lower Tc. The results in this section showed that not only Pn
had significant effect on the mixing and pressure recovery
process. The temperature To also had considerable effect on
this process.
The numerical results have shown good agreement
with experimental data in all the conditions used in this
research with some deviation in some cases. The numerical
modeling can be improved by establishing a better


4.2


4.0


~23.8


S3.6


3.4


3.2






4.6


250


200

C-
150~





50


0
0







250


200


150 C-


100


so


- -U. 1m=9 %. TPC
--U~.m=90%. TPC
- U. m=80%, TPC
- U =180%. TPC


experiment
-P~.relm=90%, noTPc
pre PEm,1=90%. TPC
--- Pr~, 11=80%, noTPC
- Pr~, Em=80%, TPc


P ,= 10MPa
To=420C Te= 50C

`L.


30 40 500


0 10 20
x [mm]
(b)






7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30-June 4, 2010


Nakagawa, M., Harada, A. and Berana, M. Analysis of
Expansion Waves Appearing in the Outlets of Two-Phase
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Paper No


correlation of the droplet size and by obtaining more
accurate motive nozzle outlet velocity. These potentials for
improvement are the subject of our ongoing research. The
numerical results also confirmed our conclusion in our
previous study that the motive nozzle and ejector becomes
more efficient as the motive inlet moves to the left of the
P-h diagram and also that the suction pressure has minimal
effect on ejector performance as compared to P,,, and T,.
The results of this numerical study aid in better
understanding the high-speed dynamics inside the CO2
two-phase ejector are shown in


CONCLUSIONS

A numerical simulation using commercial CFD code
was used in this research to model the mixing and pressure
recovery process inside a two-phase CO2 two phase ejector.
The pressure profiles obtained from CFD model have shown
good agreement with experimental data. The pressure
recovery profile and the magnitude of pressure recovery are
largely affected by the motive nozzle inlet condition. The
suction pressure yielded minimal effect on these parameters.
The effect of phase change was considerably small for the
cases used in this research. The centerline velocity of gas
and liquid phases had negligible difference particularly
during the start of pressure recovery. The energy conversion
efficiency of the motive nozzle has increased for increasing
inlet pressure and lower inlet temperature. The numerical
results confirmed our conclusion in our previous study that
this efficiency, as well as the experimental coefficient of
performance of the system, increases as the inlet condition
moves to the left of a P-h diagram. The results of this study
provide useful information on how to realize an optimum
CO2 ejector refrigeration system.


REFERENCES

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Journal of Refrigeration, Vol. 32 (6), pp.195-1202 (2009)




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