HEFAT2007 5 International Conference on Heat Transfer, Fluid Mechanics and Thermodynamics

HEFAT2007 5 International Conference on Heat Transfer, Fluid Mechanics and Thermodynamics
HEFAT2007
5th International Conference on Heat Transfer, Fluid Mechanics and Thermodynamics
Sun City, South Africa
Paper number: LA1
EXPERIMENTAL AND NUMERICAL STUDY OF BUBBLE-PUMP PERFORMANCE
OPERATED WITH A BINARY ORGANIC MIXTURE
Levy A.
Pearlstone Centre for Aeronautical Engineering Studies
Mechanical Engineering Department,
Ben-Gurion University of the Negev,
P.O. Box 653, Beer-Sheva 84105,
Israel,
E-mail: [email protected]
ABSTRACT
A two-fluid model was used to model and simulate the flow
of an organic binary mixture through a generator and a bubble
pump tube. Both thermal and non-thermal equilibrium
conditions were examined. Results obtained from a steady state
steady flow experimental system operated with R22-DMAC
(chlorodifuoromethane- N-N' dime-thylacetamide) as working
fluids were used to validate the predictions of the numerical
simulations. The predictions of the numerical simulations for
various heat sources on the flow characteristics were compared
with the experimental results. Based on the numerical
simulations detailed description of the flow characteristics
inside the generator and the bubble pump tube were obtained.
INTRODUCTION
The bubble pump is the motive force of the diffusion
absorption cycle and is a critical component of the absorption
diffusion refrigeration unit. The purpose of the bubble pump
(besides the circulation of the working fluid) is to desorb the
solute refrigerant from the solution. Therefore the efficiency of
the bubble pump is set by the amount of the refrigerant
desorbed from the solution. The performance of the diffusion
absorption cycle depends primarily on the efficiency of the
bubble pump. For that reason it was decided to investigate both
numerically and experimentally the performance of the bubble
pump for a new diffusion absorption cycle operated with R22DMAC.
When a fluid flows vertically up in a tube, which is heated
either uniformly along the entire tube length or just at the first
tube section, the flow pattern of the two-phase flow is
continuously changing along the heating section of the tube due
to mass transfer. The evaporation of the volatile phase causes a
continuous change of both the liquid and the gas flow rates.
Typical flow patterns, tube wall and fluid temperatures and the
heat transfer ranges in a vertical flow heated tube are presented
in Stephan [1] and Collier and Thome [2]. Numerical
simulation of two phase flows with heat and mass transfer
raises many different and difficult issues, from both modelling
and computational points of view. Various theoretical
approaches based on mass, momentum and energy balances can
be adopted. While considering the phases' interactions and
modes of flows, which can be found in evaporation,
condensation, absorption and desalination, the two-fluid model
seems to be the most suitable approach for describing this
process. Therefore many researchers adopted the two-fluid
model for modelling similar processes, e.g. [3-8].
NOMENCLATURE
fL0
h
mɺ
qɺ
q
F
Re
T
P
u
z
[kJ/kg K]
[kJ/kg K]
[m3K/W]
Friction coefficient
Enthalpy
Volumetric mass transfer rate
[W/m3]
[W/m3]
[N/m3]
[K]
[K]
[Pa]
[m/s]
[m]
Volumetric heat generation density
Volumetric heat transfer
Force per unit volume
Reynolds number
Temperature
Pressure
Velocity
Cartesian axis direction
Special characters
ϕ
[-]
[kg/m3]
ρ
[-]
ξ
Gas volume fraction
density
Mass fraction of the refrigerant in the solution
Subscripts
l
i
WL
rl
a
mix
g
gl
Liquid phase
Phases interface
Wall
Refrigerant in the liquid phase
Absorbent
Access property
Gas phase
Transfer property between the phases
Superscripts
*
Equilibrium condition
Figure 1 Schematic illustration of the experimental system.
Table 1: Experimental times averaged flow properties as
obtained experimentally for various heat inputs and driving
head 0.62 m..
Heat Input [Watt]
Temp. [oC]
Levy et al. [8] investigated the convective boiling of binary
organic solution in vertical tube. Based on the simplified
assumptions (One-dimensional flow; Steady state flow;
Constant heat flux at the generator; Constant tube diameter
(i.e., cross-section area); Uniform cross-section fluids
properties; Compressible gas and liquid phases; Thermal
equilibrium (i.e., both phases are at the same temperature);
Absorbent vapour in the gas phase can be neglected; Friction
force between the vapour phase and the wall per unit volume
can be neglected) balance equations for the single- and the twophase flows were developed. The developed model was used to
simulate the convective boiling of binary organic solution in
vertical tube.
Koyfman et al. [9] investigated experimentally the performance
of the bubble pump for diffusion absorption refrigeration units.
A continuous experimental system was designed, built and
operated. An experimental parametric study was conducted to
examine the influence of parameters such as heat input, motive
head, operating pressure and tube diameter on the bubble pump
performance. Based on a system analysis it was concluded that
the poor solution can be regards as in equilibrium at the bubble
pump outlet although it is not at equilibrium condition at the
generator outlet. For this reason, the assumption of thermal
equilibrium (i.e., both phases are at the same temperature) that
made by Levy et al. [8] is not justified all the way through the
generator and convective tube.
In the present study experimental and numerical study of
bubble-pump performance operated with a binary organic
mixture was conducted. The solution was composed of organic
solvent and hydrofluorocarbon refrigerant, dimethylacetamide
(DMAC) and chlorodifluorometane (R22), respectively. The
two-fluid model was used to model and simulates the flow
characteristics along the bubble-pump tube and the heating
section. In contrary to Levy et al. [8] model, non-thermal
equilibrium assumption was made. The model was used for
further system analysis and obtaining some design parameters
for the bubble-pump tube and the heating section. A continuous
experimental system was used to investigated the bubble-pump
performance and to validate the model‘s predictions.
140
160
180
200
220
240
Gen. In
27.97
29.55
30.41
28.68
29.45
30.09
Ref. H.E. in
29.98
32.82
33.84
34.48
37.42
40.61
Ref. H.E. out
25.47
26.44
26.84
25.71
26.87
28.05
P.S. H.E. in
44.33
47.78
49.57
49.23
52.87
58.37
P.S. H.E. out
35.70
40.71
42.07
41.25
43.77
45.77
Absorber in
38.81
42.51
44.20
43.90
46.72
49.11
Absorber out
35.83
38.24
39.47
38.57
40.95
42.63
Reservoir out
35.63
38.30
39.54
38.53
40.70
42.04
[10 Pa] Sys. Pressure
3.584
3.881
4.036
3.970
4.259
4.485
[10-2 Pa] Pressure diff.
51.064
50.159
49.855
49.423
49.192
50.159
[ml/min] Rich Sol. flow rate
90.90
97.32
100.50
104.46
104.58
96.00
[ml/min] Poor Sol. flow rate
75.72
83.58
85.74
88.62
87.96
77.28
-5
EXPERIMENTAL SETUP AND RESULTS
Schematic illustration of the experimental system is
presented in Fig. 1. Rich solution is pumping out from the
reservoir, cooled down and flows through a flow meter to a
generator unit where the solution gains heat. The desorption
process at the generator creates small vapour bubbles which
merge into larger bubbles. The rising bubbles form slugs that
occupy the whole cross section of the glass tube and flow with
the poor solution to the separating vessel. From the separating
vessel, the gaseous phase flows up to the gas heat exchanger
and the poor solution flows to the solution heat exchanger. Due
to the large difference of the normal boiling temperature
between the absorbent and the refrigerant (>200oC), the
presence of absorbent vapour in the gas phase may be
neglected. The cooled (not condensed) refrigerant vapour and
the poor solution enter the absorber and the absorption process
takes place. The rich solution from the absorber flows back into
the reservoir.
The heating unit, (generator) and the convective glass tube
inside diameters were 7 mm. 3/8” copper tubes were bended
into oval spirals and used as the absorber, poor solution and
refrigerant heat exchangers. T type (copper-constantan)
thermocouples, with uncertainty of ±0.3°C , were mounted
along the experimental system. A STS pressure transducer,
(ATM model - measuring range of 0-20 bar, ±0.1%FS ), was
used to measure the absolute system. The volumetric flow rates
of the poor and the rich solutions were measured by Kobold
Pelton turbine flow meters, LM model, measuring range 0.021.3 ml/min, with uncertainty of ±0.5%FS . The pressure drop
across the convective tube and the generator was measured with
a Smar differential pressure transducer, LD 301 model,
calibrated to 0-200 mbar range, with uncertainty of
±0.075%FS . Power is supplied to three 220V/100W electric
heaters by a variable transformer 0-220V and measured by a
digital wattmeter, with uncertainty of ±1W . The sensors were
connected to a personal computer using a National Instruments
DAQPad-4350 data logging system. Data logging application
was written using National Instruments LabVIEW. The system
operated continuously until a steady state was obtained. Times
averaged of the recorded properties as obtained for various heat
input are presented in Table 1.
THE HYDRODYNAMIC MODEL
As describe in the previous section, sub-cooled solution is
inserting into the generator, which located at the bottom of the
bubble-pump tube. Therefore the model should account both
the single- and two-phase flows. Mass, momentum and energy
balance equations were written for both flows. The
conservation equations for the flow of solution with constant
refrigerant concentration were solved for the single-phase flow.
Heat supplied to the solution until it reached an equilibrium
conditions, i.e., the equilibrium concentration of the refrigerant
at the solution pressure and temperature is equal to that of the
sub-cooled solution. When equilibrium achieved, refrigerant
starts to dissolve from the solution due to additional heat supply
and pressure reduction; the concentration of the refrigerant in
the solution decreases; two-phase flow is initiated while the
volume fraction of the gas phase increases. Generally, the gas
phase is composed of both absorbent and refrigerant vapours.
However, due to the large difference of the normal boiling
temperature between the absorbent and the refrigerant, the
existence of absorbent vapour in the gas phase may be
neglected, as was assumed in this study.
The developed model is based on the following
assumptions: One-dimensional flow; Steady state flow;
Constant heat flux at the generator; Constant tube diameter, i.e.,
cross-section area; Uniform cross-section fluids properties;
Compressible gas and liquid phases; Absorbent vapour in the
gas phase can be neglected; Friction force between the gas
phase and the wall per unit volume can be neglected. Finally
thermal and non-thermal equilibrium conditions between the
phases were examined (i.e., both phases having the same
temperature or different temperatures, respectively).
Based on these simplified assumptions balance equations
for the single- and the two-phase flows were written.
(
Single-Phase Flow Governing Equations ξ * > ξ
)
Solution mass, momentum and energy balances
∂
( ρl ul ) = 0
∂z
(1)
(2)

u2  
∂ 
 ρ l ul  hl + l   = qɺ
∂z 
2 

(3)
)
The solution's enthalpy, hl ( ≡ ξ hrl + (1 − ξ ) ha +△hmix ) , was
calculated as a function of the composition, the enthalpy of the
refrigerant in the liquid phase, hrl = f ( T ) , the enthalpy of the
ha = f (T )
and
solution's
excess
enthalpy,
△hmix = f (T , ξ ) [10]; The solution density was calculated as
function of the composition, refrigerant and absorbent densities
[11]. The single-phase balance equations were solved
numerically. The refrigerant concentration kept constant and
when it reached the value of the equilibrium concentration,
ξ * = f (T , P ) , the single-phase flow solver stopped and the
two-phases flow solver was initiated.
Two-Phase Flow Governing Equations
Total, vapour phase and refrigerant mass balances
∂
(1 − ϕ ) ρl ul + ϕρ g ug = 0
∂z
(4)
∂
ϕρ g u g = mɺ
∂z
(5)
(
(
)
)
∂
(1 − ϕ ) ξρl ul + ϕρ g u g = 0
∂z
(
)
(6)
Vapour phase and total momentum balances
∂
∂P
ɺ i − Fgl
ϕρ g u g 2 + ϕ
= mu
∂z
∂z
(7)
∂
∂
∂P
ϕρ g u g 2 + ( (1 − ϕ ) ρl ul ul ) +
= − FWL
∂z
∂z
∂z
(8)
(
(
)
)
Total energy balance

u2   ∂ 

u2  
∂ 
 ϕρ g u g  hg + g   +  (1 − ϕ ) ρl ul  hl + l   = qɺ (9)

∂z 
2   ∂z 
2  




As mentioned above, in the present study both thermal and
non-thermal equilibrium conditions were examined. When
thermal equilibrium condition was examined it was assumed
that both phases have the same temperature and therefore only
the total energy balance equation was used. However for the
non-thermal equilibrium condition (i.e. both phases have
different temperatures), the gas phase energy balance equation
was added.
∂
ɺ i + qgl
ϕρ g u g hg = mh
∂z
(
∂
∂P
ρl ul 2 +
= − FWL
∂z
∂z
(
absorbent,
)
(10)
The vapour phase enthalpy was taken as the enthalpy of the
refrigerant as function of pressure and temperature. Three
additional models for calculating the wall and vapour-liquid
friction forces and the heat transfer between the phases (for
non-thermal equilibrium condition) were introduced in order to
solve the complete model. The additional models that adopted
here were used and recommended by Richter [3] & Yang and
Zhang [7]. The liquid wall friction force introduce from
Martinelli & Nelson [12].
FWL
(ϕρ u + (1 − ϕ ) ρ u )
2 (ϕρ + (1 − ϕ ) ρ ) D
g
g
g
f Lo = ( 0.79ℓn Re Lo − 1.64 )
Table 2: Models compressions with experimental data.
2
l l
l
(11)
pipe
−2
The interfacial force between the phase, Fgl, for bubble
(ϕ < 0.3) and annular (ϕ > 0.8) flow regimes assumed to be
known (For more details see [3], [7] and [8]).
Fgl =
3 Cd*
3
ϕ (1 − ϕ ) ρl u g − ul ug − ul
4 d
(
(
)
)
C
Fgl = 3 d ϕ ρ g u g − ul ug − ul
D pipe
(ϕ < 0.3)
(ϕ > 0.8 )
(12)
The drag coefficient for bubbly flow, Cd = Cd (1 − ϕ ) ,
reflects the influence of the bubbles on each other and therefore
depends on the gas phase volume fraction [13].
*
 24
(1 + 0.15 Re0.687
) Red ≤ 1000
d

Cd =  Re d

0.44
Re d > 1000

(ϕ < 0.3)
Cd = 0.005 (1 + 75 (1 − ϕ ) )
(ϕ > 0.8 )
−4.7
(13)
The momentum balances of both phases are valid for all
flow regimes that might be observed in the bubble pump tube,
however the interfacial forces that are known are restricted to
bubbly and annular flow regimes, where the gas volume
fraction is below 0.3 or above 0.8, respectively. In the present
study it was assumed that the drag force in between bubbly and
annular flow regimes (i.e., in the intermediate flow regimes:
plug flow, churn flow and wispy-annular flow) can be linearly
interpolated with respect to the gas phase volume fraction.
Similar assumption was used for calculating the heat transfer
coefficient between the phases when non-thermal equilibrium
conditions were examined.
The numerical solution of the above balance equations was
obtained using Gear's fifth order BDF method, which is
available in the IMSL library. The simulation starts at ahead of
the boiling chamber and stopped at the end of the bubble pump
tube (i.e., separation vessel).
Temp.
Pressure drop [10-2 Pa] Poor solution volume Poor Solution
[oC]
flow rate [ml/min]
1.75
 1− x 
= f Lo 

 1−ϕ 
Heat Input [Watt]
Experimental SD%
Prediction of TEM
TEM Error
160
0.05%
180
0.05%
200
0.06%
220
0.05%
240
0.09%
52.08
56.03
59.50
59.96
64.80
72.98
-2.44%
-2.57%
-3.08%
-3.33%
-3.66%
-4.41%
52.08
56.03
59.50
59.96
64.80
72.98
NTEM Error
-2.44%
-2.57%
-3.08%
-3.33%
-3.66%
-4.41%
Experimental SD%
1.17%
1.33%
2.83%
2.87%
1.24%
6.25%
Prediction of TEM
78.09
82.93
84.32
86.56
85.22
75.05
-3.13%
0.77%
1.66%
2.33%
3.11%
2.88%
Prediction of NTEM
TEM Error
78.09
82.93
84.32
86.56
85.22
75.05
NTEM Error
-3.13%
0.77%
1.66%
2.33%
3.11%
2.88%
Experimental SD%
2.13%
1.97%
2.18%
2.78%
3.02%
4.11%
Prediction of TEM
50.98
50.94
50.18
49.72
49.13
49.64
TEM Error
0.16%
-1.56%
-0.65%
-0.60%
0.13%
1.04%
Prediction of NTEM
Prediction of NTEM
50.98
50.94
50.18
49.72
49.13
49.64
NTEM Error
0.16%
-1.56%
-0.65%
-0.60%
0.13%
1.04%
continuous liquid properties were used, and a very small gas
mass flow rate per unit area was assumed (10-8 kg/m2). To
conserve the total mass balance, the mass flow rate of the gas
phase was subtracted from that of the liquid phase and then the
gas volume fraction was calculated. By integrating the balance
equations along the convective tube, the outlet conditions of the
flow at the separation vessel were predicted.
Three additional assumptions have been made in order to
validate the numerical simulations predictions. The first, the
poor solution temperature at the separation vessel is similar to
the one measured at the poor solution heat exchanger inlet (see
Fig. 1). The second, the pressure drop across the separation
vessel is negligibly small and therefore the predicted pressure
drop is equal to the measured one. The third, the predicted
volume flow rate of the poor solution is equal to the measured
solution's volume flow rate at the outlet of the poor solution
heat exchanger. This assumption can be justified by an efficient
separation process, i.e., the separated gas can not flow into the
solution heat exchanger due to a liquid trap. Based on these
assumptions the predictions of the simulations could compare
with the experimental data. Table 2 present comparisons
between the predictions of the Thermal Equilibrium Model
(TEM) and the Non-Thermal Equilibrium Model (NTEM) and
the experimental data for the poor solution temperature, volume
flow rate and the pressure drop. An error is defined by
Error [%] =
MODELS VALIDATION AND COMPARISON
The models were solved numerically for simulating the
experiments with various heat inputs, rich solution mass flow
rate and operating pressure, in which their times averaged
properties are presented in Table 1. The following inlet
conditions were specified: solution's temperature, operating
pressure, concentration and volumetric flow rate. The inlet
concentration was calculated, based on vapour liquid
equilibrium (VLE) assumption at the reservoir, as a function of
the reservoir temperature and the system pressure. The solution
inlet velocity was calculated from the volumetric flow rate. In
the transition between the single- and the two-phase models,
140
0.03%
Experimental Value - Predicted Value
Experimental Value
(14)
The validity of the TEM and the NTEM and their
assumptions can be clearly seen in Table 2. The predicted
pressure drops along the generator and the bubble pump tube
varied between 1.1% to -1.6% from the experimental values.
As can also be seen, the deviations of the predicted pressure
drops from the experimental values are less that the
experimental standard deviation values. The predicted poor
solution flow rates varied between 3.2% to -3.2% from the
experimental values, while the experimental standard
deviations were up to 6.25%. Both models over predicted the
poor solution temperature at the bubble pump outlet by up to
4.5%. This might be due to the assumption that the poor
0.8
0.7
0.6
φ,ξ
0.5
0.4
0.3
0.2
(a)
0.1
0
0.14
0.16
0.18
0.2
0.22
0.24
0.26
0.28
0.3
z [m]
1
u , u [m/s]
0.8
l
0.6
g
0.4
0.2
(b)
0
0.14
0.16
0.18
0.2
0.22
0.24
0.26
0.28
0.3
z [m]
65
60
50
g
l
T , T [oC]
55
45
40
35
30
(c)
0.14
0.16
0.18
0.2
0.22
0.24
0.26
0.28
0.3
z [m]
410
409
P [kPa]
solution temperature at the separation vessel is similar to that
measured at the poor solution heat exchanger inlet. Therefore,
the difference between the predicted values at the bubble pump
outlet and the measured at the poor solution heat exchanger
inlet can be justified by heat loses.
The predictions of the TEM and the NTEM show similar
values at the bubble pump outlet. This results form the
equilibrium conditions achieved in the bubble pump tube.
Figure 2 presents comparison between the predictions of the
TEM (black lines) and the NREM (blue lines) for the gas
volume fraction, solution concentration, phases' velocities and
temperatures and pressure along the heating section and the part
of the bubble pump tube as obtained by the numerical
simulations for 180W generator (see Table 1). The four vertical
dashed lines represent locations within the heating section. The
first and the last mark the location of the heating section while
the second and the third mark the transition from sub-cooled
flow to bubble flow and then to plug, churn or wispy-annular
flow regimes, respectively.
As can be seen in Figure 2, sub-cooled solution is enter the
vertical tube, the pressure is being drop due to gravity and wall
friction until it enters to the heating section. The solution is
being heat rapidly until the solution reached the equilibrium
temperature. At this point (second vertical line), further heating
results in gas generation by desorbing of refrigerate from the
solution. Bubbles are formed and due to densities difference
(gravity), their velocity increases rapidly towards their terminal
velocity. Due to the increase in the average diameter of the
bubbles and the increase of gas volume fraction, the velocity of
the bubbles then starts to decrease (Fig. 2b). The refrigerant
desorption from the solution causes a decrease in the rate of
temperature and pressure changes. The gas phase is
accelerating and drags the liquid upwards. Additional gas is
being separated from the solution due to both pressure
reduction (flashing) and temperature increases (heat supplied).
When the two-phase flow leaves the heating section, the
pressure is continuously dropped due to friction and gravity.
This results in minor changes in phases' velocities and
temperature. In the TEM, the small reduction of the pressure
cause to a minor decrease in both the liquid and the gas
temperatures. In the NTEM, the gas phase temperature
increases due to heat exchange with the liquid phase, until it
reaches thermal equilibrium condition. From that point onward
a minor decrease in both the liquid and the gas temperatures
may be observed due to the small pressure reduction while the
solution concentration and gas volume fraction practically
remains constant.
The main difference between the TEM and the NTEM can be
seen in Figure 2c. In the NTEM, the gas phase temperature (the
dashed blue line) is lower than the equilibrium temperature.
However, due to high heat and mass transfer between the
phases in the bubbly flow regime, this temperature difference is
negligible. Except from the lower gas phase temperature at the
generator outlet, there is no much difference between the
models. Therefore the TEM model, in which used less semiempirical models for calculating the heat transfer between the
phases might be preferable for further analysis and system
design.
408
407
406
(d)
405
0.14
0.16
0.18
0.2
0.22
0.24
0.26
0.28
0.3
z [m]
Figure 2 Comparison between the predictions of the TEM
(black lines) and the NTEM (blue lines) as obtained for
180Watt heat input. (a) gas volume fraction (solid line) &
solution concentration (dashed line), (b) gas and liquid phases'
velocities (solid & dashed lines, respectively), (c) liquid and
gas phases’ temperatures (solid & dashed lines, respectively)
and (d) pressure along the vertical tube.
THE INFLUENCE OF THE HEAT INPUT
The influence of the heat input to the generator on the
bubble pump performance was investigated. It was obtained
that for constant driving head increasing the heat input results
in higher rich and poor solutions flow rates, higher outlet
temperature and lower pressure drop across the bubble pump
and the bubble pump tube. The reduction of the pressure drop
resulted from the higher refrigerant volume fraction in the
bubble pump tube for higher heat input. It should be noted that
although the poor solution flow rate increases, more refrigerant
dissolved from the solution for higher heat input. Therefore the
pumping ratio (defined as the ratio between the poor solution
and the refrigerant mass flow rates) decreases as the heat input
increases. The influence of the driving head on the bubble
pump performances was also investigated. It was obtained that
increasing the driving head increases both the pressure drop and
the poor solution flow rate.
On the basis of the numerical simulations, a detailed
description of the flow characteristics inside the generator and
the bubble pump tube was obtained. The influence of the heat
source and driving head on the flow characteristics was
examined both numerically and experimentally.
The comparison between the TEM and the NTEM shows that
there is no much difference between them if one looks at the
overall performances of the unit. The NTEM relays on more
semi-empirical correlations and assumption for calculating the
heat transfer between the phases. The inaccuracies of these
correlations and additional models may lead to larger
discrepancy between the prediction of the numerical
simulations and the experimental data. Nevertheless the overall
energy balance was conserved, and therefore there was no
difference between the predictions of the TEM and the NTEM
at the bubble pump outlet. Therefore, for simplicity the TEM
should be used for further analysis and system design.
THE INFLUENCE OF THE OPERATING PRESSURE
The influence of the operating pressure on the flow
characteristics was examined numerically. The predictions were
obtained for the inlet conditions specified in Table 1 with
200W input and various operating pressures. These predictions
could not be validated since in a closed continuous system,
such as ours, the operating pressure and inlet temperature
cannot be controlled for a given heat input. Since the rich
solution in the reservoir (see Fig. 1) is in equilibrium condition,
higher refrigerant concentration obtained when the operating
pressure is higher. For the same heat input, more refrigerant
desorbed from the solution, when solution with higher
refrigerant concentration flow through the generator and the
bubble pump tube. Nevertheless, the gas volume fraction was
decreased. This resulted from the hydrodynamic model which
considered compressible gas and liquid phases. Denser gas and
liquid phases result in lower phases' velocities. The influence of
the operating pressure on the solution temperature and pressure
drop was negligibly small.
REFERENCES
CONCLUSION
Numerical and experimental study of bubble pump
performance was conducted. A continuous experimental system
was operated to validate the predictions of the numerical
simulations. A two-fluid model was used to model flows in
both thermal and non-thermal equilibrium conditions. The
models were solved numerically to simulate the flow and to
obtain the characteristic flow profiles along the generator and
the bubble pump tube. The predictions of the numerical
simulations were validated experimentally.
The assumption that due to the large difference between the
normal boiling temperature of the absorbent and that of the
refrigerant (more than 200°C), the presence of absorbent
vapour in the gas phase could be neglected was justified by the
good agreement between the predictions of both models and the
experimental data. However, this assumption limiting the
models to simulating flow boiling of binary mixtures for which
there is a large difference of normal boiling temperature
between absorbent and refrigerant.
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