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Dynamic simulation of an aqua-ammonia absorption cooling system with refrigerant storage
Energy Convers.Mgmt Vol. 32, No. 3, pp. 197-206, 1991
0196-8904/91 $3.00+ 0.00 Copyright © 1991 Pergamon Press pie
Printed in Great Britain. All rights reserved
D Y N A M I C SIMULATION OF AN A Q U A - A M M O N I A ABSORPTION COOLING SYSTEM WITH REFRIGERANT STORAGE S. C. KAUSHIK, S. K. RAO and RAJESH KUMARI Centre of Energy Studies, Indian Institute of Technology, Hauz Khas, New Delhi-110 016, India
(Received 18 September 1989; receivedfor publication 30 August 1990) A b s t r a c t - - A dynamic computer simulation model of a solar powered a q u a - a m m o n i a absorption cooling system with refrigerant storage is presented. The model requires a known solution flow-rate as well as average hourly weather data and the cooling load pattern as input parameters and is capable of predicting various heat flows, temperatures, concentrations, mass flows and mass stored in the absorber/condenser stores. The design parameters of the system components are evaluated at rated conditions using the steady state model. Test simulation of the dynamic model is carried out with the cooling load corresponding to a typical cold storage house on a hot summer day. It is found that the refrigerant stored during excess generation hours m a y not be sufficient to meet the cooling load during off-generation hours, and hence, provision to evaluate the auxiliary refrigerant mass required and the fraction of the load met by solar energy is included in the simulation model. Thus, the concept of refrigerant storage within the cycle is feasible, but it has to be integrated with other storage options for absorption cooling systems. Absorption cooling Dynamic simulation ous space conditioning
Aqua-ammonia
Refrigerant storage
NOMENCLATURE A = Cp = E = h = ~r = MPR = P = Q = t = U = u = v= x =
Area (m 2) Specific heat of water/air (kJ/kg°C) Internal energy (kJ/kg) Enthalpy of fluid (k J) Mass flow-rate of refrigerant/absorbent (kg/s) Mass pumping ratio Pressure (kPa) Heat transfer rate (W/m:) Temperature (°C) Heat transfer coet~cient (kW/m 2 °C) Specific internal energy (kJ/kg) Specific volume (m3/kg) Concentration of refrigerant in solution (kg/kg)
Subscripts a = A= AS = C = CS = E = G = i= 1= o= p= r= R = S= s= V= W = w= I, II = 1-17 =
Air (ambient/cooling space for low temperature heat supply) Absorber Absorber store Condenser Condenser store Evaporator Generator Inlet collector fluid Liquid phase Outlet collector fluid Collector parameter Refrigerant Rectifier Solar radiation Strong solution (rich in refrigerant) Vapour phase Water Weak solution (poor in refrigerant) Type of heat exchanger (preheater/subeooler) State points
Greek letters 0 = Time (h) = Absorptivity ECM 32/3---A
197
Continu-
198
K A U S H I K et al.:
A Q U A - A M M O N I A ABSORPTION COOLING SYSTEM
z = Transmittance E = Nonequilibrium approach factor r / = Effectiveness of heat exchangers
INTRODUCTION In solar operation of absorption space conditioning [1-3] systems, heating/cooling is accomplished by solar energy during sunshine hours and by an auxiliary source of energy during off-sunshine hours. Since the supply of solar energy and the demand of heating/cooling load are time dependent with certain phase lags, energy storage becomes essential for the efficient use of these absorption cycle systems. Several types of energy storage, viz. hot water storage or cold water storage etc., are possible in conjunction with solar operated absorption systems, but these storage systems have low storage capacity per unit volume in addition to higher thermal loss to ambient and are bulky in nature, and hence, these storage systems can be used for domestic size solar space conditioning. The use of phase change storage media can reduce the storage volume, but their performance gets degraded by 'repeated usage. Recently, a novel concept of solar energy storage by refrigerant storage (utilizing the latent heat of evaporation within the absorption cycle itself) has been proposed by a number of workers in the literature [4-6]. Thereafter, several authors [7-10] have studied the refrigerant storage concept for solar cooling. The working fluid pair in their study was LiBr-H20. However, it is also worth studying the NH3-H20 absorption cooling system, since it has some potential advantages (like no-crystallization etc.) over the LiBr-H20 system [11, 12]. In the present communication, the authors have carried out a dynamic simulation of an NH3-H~O absorption cooling system with refrigerant storage to study the time dependent system performance for matching of the cooling load corresponding to a typical hot summer day. AN
ABSORPTION
AIR
CONDITIONER
WITH
REFRIGERANT
STORAGE
The system considered is shown in Fig. 1. It includes the generator, rectifier, condenser, evaporator and absorber along with preheating/subcooling heat exchangers, pump and pressure ] I "1"C2 I
I-
I
~ /
Condenser
~- I
I I t 12
TRSZ ~ l _ ,, C..T. store~"~'~cs _ i
Refrlgeront
13
r~s~
exchonoer
19
t,i l!
1
Fig. 1. An absorption airconditioner with refrigerant storage.
I I I I I J
KAUSHIK et al.: AQUA-AMMONIAABSORPTION COOLING SYSTEM
199
reducing values. In refrigerant storage absorption cycle systems, a storage volume is provided in association with the condenser to store the refrigerant during the hours of excess generation (i.e. hours of high insolation). The stored refrigerant is released in a manner to meet the required load during off-generation hours. Storage is also needed in the absorber to accommodate not only the refrigerant but also sufficient absorbent to keep the concentration within allowable limits. The condenser store is assumed to have a regulating value at its outlet to control the refrigerant flow. Heat input to the generator is supplied through a flat plate solar collector. The heat of condensation at the condenser and the heat of solution at the absorber are rejected to the atmosphere by a cooling tower, from which water circulates through the absorber store, absorber condenser store and condenser. Room air is circulated through the evaporator at constant inlet temperature. SYSTEM MODEL AND BASIC EQUATIONS In addition to the usual assumptions [11] made in modelling any solar powered absorption cycle system, the following are the two extra assumptions made for modelling the solar powered absorption cooling system with refrigerant and absorber stores. (i) The storage rate remains constant over the time interval under consideration, and the total storage for the interval can be found by first order approximation, (ii) The mass of refrigerant at any time in components other than the stores is neglible. Modelling of different components of the system is discussed next. Solar collector
A fiat plate collector supplies the heat input to the generator. The flat plate collector is modelled according to the collector performance equation [12] given by Qp = Fp {(ez)Q~(0) - Up[ti - ta(O)lAp.
(1)
This Qp is also equal to the heat taken by the circulating heat transfer fluid Qp = 3;/pCw (to - ti).
(2)
The useful heat from the collector is transfered to the generator through the heat exchanger relation. Therefore, we have
Qp=QG=(UA)c
[(to -- tc) -- (ti -- tc)] lnr(to_ tc) 1
(3)
/(ti- tc)_] Generator
The generator is assumed to be a heat exchanger with the collector heat transfer fluid flowing on one side of the heat exchanger and a two-phase flow on the other side. Since a larger part of the generator heat is utilized for boiling of the generator solution, the lower side of the heat exchanger is assumed to be at outlet temperature. The residential time in the generator is too small for equilibrium conditions to be set up. Therefore, an equilibrium approaching factor E is defined as
Xs-Xw E
Xs- x*
(a)
the ratio of the differences in concentrations of the ammonia solutions leaving the generator and absorber. E = 1 corresponds to the full equilibrium state. Here X* is the equilibrium concentration of the solution in the generator at the pressure and temperature of the generator. The basic energy balance equation for the generator is {~G "JI- 2~f,h 5 -3t- .~rl0hl0 = J~rrh 6 --[- 2~/9h 9
where the state points correspond to Fig. 1.
(5)
200
KAUSHIK et
al.:
AQUA-AMMONIA ABSORPTION COOLING SYSTEM
Rectifier The rectifier is also a heat exchanger. It is assumed that the temperature throughout the rectifier is the same. The following are the energy balances around the rectifier
-~fgh9 =/~fl0hl0 q" ~fll hll + QR
(6a)
QR = A:/'4h, - A:/'5h5 = (UA)R [(tz -- ts) -- (tR -- t,)] l n ( t R - ts'~ \tR - t4}
(6b)
In the above equations, QR is the heat evolved at the rectifier. This heat is normally fed to the strong solution before it enters the generator to reduce the generator heat input requirement. Condenser The condenser is also modelled as a heat exchanger. However, the process in condenser heat transfer is different from that in the other heat exchangers of the system because of the condensation of the refrigerant. The UA product of the condenser heat exchanger becomes a function of the refrigerant flow-rate as the heat transfer due to condensation varies over a wide range with flow-rate and, hence, with the thickness of the condensate. Its UA product varies with the flow-rate of the refrigerant since the thickness of the condensate and, hence, the heat transfer rate vary with flow-rate. An equation developed by Grassie and Sheridan [7] is used to determine the actual UA product of the condenser 3.77 (UA)c = 2.4 + 5.5 MIt 3 (UA)r~
(7a)
where (UA)s is the nominal rating of the condenser. The above equation holds good only for small variations over 3.0 kW/s of (UA)N. For the condenser, one can write the following energy balance equations. A;/llh~ = M12h12 + Qc
(7b)
Q¢ = AlaCw(t¢2 - t c l ) = (UA)c [(tc- tel) - (to - to2)] ln(t¢-- tcl~ \t¢ - tc~}
(7c)
Condenser store Mass and energy balances on the condenser store are given by two first order differential equations dMcs = A;/i~ _ A:/l3 dO
dEcs
Ml~hl2 = Might3 + ~
+
(8a)
Qcs
dEcs dMcs . dues dO = dO ucs + M c s dO
(8b)
(8c)
where u is the specific internal energy and is obtained from the state equation (u = h - p v , v is the specific volume) and ....
[(tcs -- tcsl) -- (tcs -- tcs2)]
(8d)
K AU S H I K et al.: A Q U A - A M M O N I A ABSORPTION COOLING SYSTEM
201
The concentration of the refrigerant leaving the condenser store is calculated from the mass and ammonia balance conditions and is given by x5 =
M . 12,zx v . 12 ~ J i .CS - I X'~f 1 T ,trJ
(8e)
M'~2 + M~:s i
Subcooler
The subcooler is a liquid-vapour heat exchanger and is characterized by the effectiveness given by hi6 - hi8 r/n = hi 6 - his,13
(9a)
where hts, l 3 is the enthalpy of the vapour at the concentration of state 18 and temperature of state 13. The energy balance around the subcooler is (9b)
~1f13 hi3 "l-/~fl6hl6 =/~f14h14 -q- 1~flshls. Evaporator
The evaporator is modelled as a heat exchanger with the following assumptions. (i) Purge liquid leaves the evaporator at room air temperature rather than the evaporator temperature, (ii) An expansion valve can be designed to allow the required flow from the condenser store, and the refrigerant is evaporated at constant temperature corresponding to the low side pressure set by the equilibrium conditions in the absorber. The energy balance equations for the evaporator are
QE + Mishls = Ml6hl6 ']- Mi7h17
(10a)
.... [(tE, -- tE) - (tE2 -- tE)] Oe=J~IECw(tEI--tE2)=IU/I,E ln---f-~-E7 ~- )---tE ~
(10b)
L(tE2-
tE)J
Absorber
Equilibrium conditions are assumed to exist in the absorber heat exchanger. The low side pressure is determined by the strong solution concentration at state 1 and the temperature of the absorber. The basic equations for the absorber are QA + ,~/l hi = ~;18h8 + -~t19h19
QA = M , Ca(tm -- tAI) = (UA)A
[(tA- tA, ) - ( t A - tA2)] In[/A/- tAl/"]
(1 la) (1 lb)
LtA -- tA2_] Absorber store
The absorber store is similar to the condenser store. The mass and energy balances in the absorber store are given by two first order differential equations which can be solved by first order approximation, dMAs = ,~t t - M 2 dO
(12a)
.~rl hi = -,f/2h2 --t dEAs dO
(12b)
dEAs dMns dO
-
dO
dU,o
UAS + ~
OU
MAs.
(12c)
202
KAUSHIK et al.:
AQUA-AMMONIA ABSORPTION COOLING SYSTEM
The concentration of the strong solution leaving the absorber store is determined by mass and ammonia balances by assuming that the mass accumulated in the components other than the condenser and absorber stores is negligible, i.e. Mi = MAs + Mcs Xs =
(13a)
MiXi - Mcs X,3
(13b)
nAs
The solution heat exchanger The solution heat exchanger is modelled as a liquid-liquid heat exchanger whose effectiveness is given by h 6 -- h 7
rh = h6 - h7,3
(14a)
where h7,3 is the enthalpy determined at X7 and t3. The energy balance around this heat exchanger can be written as (14b)
M3h3 @/~/'6h6 = J~4h4 d-/l:/Th 7 .
Cooling load A cold storage house cooling load, varying sinusoidally, with a peak variation of 3 kW about a mean load of 7 kW is considered. This load corresponds to the house having area of approx. 250m 2 (living space) and having a heat loss coefficient of UB= 1.5W/m2°C in Brisbane (Australia) [7] and is shown in Fig. 2. The analytical representation of the cooling load considered is QEDES. =
7.0 - 3.0 cos[n (0 - 2)/12] in kilowatts.
(15)
Cooling tower model In the cooling tower model, it is assumed that the outlet water temperature can be expressed empirically as a quadratic function of both the inlet water temperature and ambient wet-bulb temperature. " ,2 . t 2 tw2 = K, + K:'twb + K, + t~b + K,t~, + K,t~, + Krtwb" tw, + K, t2wb't,i + Kstwb t2' _L T k"*''wb w,"
32
lO
3O
\\
28
1.0
3
o. o.e
24
2
22
20
)
I 4
f
I 8
I 12
16
20
Time (h)
Fig. 2. Hourly solar radiation, ambient air temperature and cooling load.
0.2
24
~
KAUSHIK et al.: AQUA-AMMONIA ABSORPTION COOLING SYSTEM
203
The constants K l - g 9 are evaluated by determining the cold water temperature, tw2 for three hot water temperatures, tw,, at each one of three wet bulb temperatures, twb. There are, thus, nine values o f the function tw2(twb -- twt) at independent values o f the variables twb and twt; hence, the K~ - K 9 can be calculated by a simple inversion matrix. NUMERICAL
COMPUTATION
AND S E L E C T I O N
OF P A R A M E T E R S
It is a rather difficult task to model the dynamic operation of an a q u a - a m m o n i a absorption cycle with all its ancillary equipment in a single stage. The calculation procedure for steady-state operation o f a simple absorption cycle, given the c o m p o n e n t temperatures and refrigerant mass flow-rate, is well established and is similar to that for any t h e r m o d y n a m i c cycle. An extension o f this steady-state analysis for the refined absorption cycle with integral refrigerant storage has been m a d e using appropriate assumptions in the present communication. The first step towards modelling the system is to identify the possible input conditions of the cycle. The input variables in the present case are assumed to be the cooling load, solar radiation and ambient conditions along with the refrigerant concentration and the solution mass flow-rate. The effectiveness of the solution heat exchanger and the subcooler and nonequilibrium a p p r o a c h factor in the generator are also assumed to be known. Initial estimates of the parameters can be m a d e on the basis of some assumed steady-state operations; however, the machine will rarely operate at such steady state conditions, During the day, considerably m o r e refrigerant will be generated than is immediately required to meet the cooling load, whereas overnight, the absorber and e v a p o r a t o r must still operate to supply the cooling load, yet no generation will take place. F o r design calculations, fixed cold storage r o o m air temperture ( ~ 10°C) and fixed supply inlet water temperature (-,~ 20°C) were chosen, the (UA) values o f the various heat exchangers were chosen to give a log mean temperature difference o f 6.6°C when transferring its m e a n load to the cooling water/air, and later, all other UA values and mass flows were adjusted to give consistent results. The values of the p a r a m e t e r s chosen in the simulation model are as shown in Table 1.
Collector parameters Ap = 9 0 m 2, (~r) = 0.9, Fp = 0.9, Ue = 4.4 W / m 2 °C, ~t;/p = A;/Gw.
Building parameters AB = 250 m 2, UB = 1.5 W / m 2 °C.
Cooling tower constants (following Grassie [7]) Kt = -0.21285547E3,
/(2 = 0.50199585E1,
K3 = 0.22859833E - 1,
K4 = 0.72521973E1,
K 5 = - 0 . 4 3 2 7 5 3 5 6 E - 1,
K6 = -0.15433693,
/(7 = 0.92089176E - 3,
/(8 = 0.10446012E - 2,
K9 = 0.64719352E - 5.
Auxiliary mass needed = 100 kg (after midnight 2.00 h). DISCUSSION
OF RESULTS
T o have a numerical appreciation of the concept of refrigerant storage for solar air conditioning, a steady-state simulation based on mass and energy balance equations of different c o m p o n e n t s of Table 1. Selection of parameters Parameter value Parameter value Parameter value (UA)G = 5.625 kW/°C E = 0.8 ~/Gw = 0.95 kg/s (UA)AB= 3.8 kW/°C rh = 0.8 M.A~ = 1.115 kg/s (UA)c = 3.8 kW/°C ~/n= 0.7 Mew = 1.115 kg/s (UA)R = 1.4 kW/°C MABi = 850 kg MEA= 1.0 kg/s (UA)e = 0.75 kW/°C Xi = 0.5 h':/Asw = 1.115 kg/s (UA)AS= 1.2 kW/°C 5;/s = 0.0015 kg/s m2 ~tcs w = I. 115 kg/s (UA)cs = 0.3 kW/°C Cpw= 4.198 kJ/kg°C CpA= 1.005 kJ/kg°C
204
K A U S H I K et al.:
A Q U A - A M M O N I A A B S O R P T I O N C O O L I N G SYSTEM
the system (excluding stores) has been developed to predict the approximate values of the design parameters. The thermodynamic properties of aqua-ammonia are used in the form of state equations expressed by earlier authors [12]. The input parameters for this model include input heat supply, constant room/ambient air temperatures, mass flow-rate of strong solutions, UA values of the component heat exchangers, effectiveness of preheater and subcooler heat exchangers, mass flow-rates of heat transfer fluids, non-equilibrium approach factor in the generator (generator is assumed to be under nonequilibrium state for the given mass-flow of the strong solution) and the design parameters of the flat plate collector. Based on the steady-state analysis, the model has then been extended for dynamic simulation of an absorption air conditioner (following Kaushik et al. [11]) including the refrigerant and absorbent stores within the cycle. To study the feasibility of refrigerant storage within the aqua-ammonia absorption cycle, a test simulation is carried out for a typical hot summer day in Brisbane (Australia). The typical hourly data for the same day for solar radiation, wet and dry bulb ambient air temperatures and the cooling load of a building are shown in Fig. 2. The dynamic simulation model uses hourly solar radiation, ambient (wet/dry) air temperatures and the desired cooling load as the input hourly variables and predicts the system component temperatures, various heat flow rates, mass flow-rates, solution concentrations and the mass stored in the condenser/absorber. The matching of the cooling load with the available cooling produced is also evaluated to determine the auxiliary refrigerant mass needed to operate the system during the off-generation period by careful control of the refrigerant flow from the store/auxiliary refrigerant mass source. Figures 3 and 4 show the hourly variation of the system component temperatures and heat transfer rates, respectively, for matching the typical cooling load corresponding to a hot summer day for which the hourly input data have been used in the simulation model. The following observations can be made from these figures: (i) For a maximum dry bulb temperature of 33°C and wet bulb temperatures of 25°C, the cooling temperature of 10°C can be achieved. The evaporator temperature is a minimum when the availability of solar radiation and dry and wet bulb temperatures are maximum. (ii) The maximum generator temperature is about 130°C, while the condenser temperature is about 42°C and the absorber temperature is about 34°C.
-60
160
'°t
50
140
4o
n
TG,To :~
120
30
--~TAB
l
,o! 80
T
60
0
I
I
I
I
I
I
I
3
6
9
12
15
le
21
Time (h) Fig. 3. Hourly variation o f component temperatures.
:0 24
K A U S H I K et al.:
A Q U A - A M M O N I A A B S O R P T I O N C O O L I N G SYSTEM
205
30
50
25 4.0
ca
E QA8
QG
Qc
15 QE
30
~. p
lo
Q) -r
•
"1"
20
10
O
3
6
9
12
I5
18
Zl
24
Time (h) Fig. 4. Hourly variation of component heat transfer rates.
(iii) The generator heat input, rectifier heat and condenser heat rejection are finite during sunshine hours, while the absorber heat and the evaporator cooling load are finite for all hours. It is also seen that QABis a maximum around 17.00 h, while QE is desired to be a maximum around 14.00 h. (iv) The generator and rectifier temperatures during the non-generation period are just the extra-plotted dashed curves, as there is no justification to evaluate these during these hours. It must be mentioned here that, in order to maintain a constant building temperature, the cooling available from the absorption air conditioner should match the building cooling load. When the cooling available is more than the load, the excess refrigerant generated is stored in the condenser store, i.e. the cooling rate is reduced to that fraction of the time step which would make the cooling produced equal to the cooling required for the total time step. For this portion of the time step, the machine used whatever solar energy was available to generate and store refrigerant. When the cooling rate becomes less than the load, the stored refrigerant is released to meet the load. If the refrigerant available, both from the store and from what is then being generated, is insufficient to sustain the load, the cooling rate is controlled by the maximum refrigerant flow-rate available from the store, and also auxiliary energy could be used to meet the load. However, when the refrigerant mass stored in the condenser store is reduced to zero, then auxiliary mass of refrigerant is added in the condenser store to maintain the system operation. The hourly variations of the mass stored in the absorber and condenser stores as a function of time are shown in Table 2. It is also possible to evaluate the fraction of the cooling load met by solar power and the auxiliary mass of refrigerant required for continuous operation of the system. It can be seen from this table that, from 8.00 to 2.00 h, the total mass (MAS+ Mcs) is equal to mi = 850 kg. However when Mcs is reduced to zero, it is assumed that the extra refrigerant mass of 50 kg is added to maintain the system operation. This table also shows the hourly variation of the refrigerant mass required to match the load and the refrigerant mass available during the system operation. Thus, the size of the storage tank can be determined from these considerations. Thus, the concept of refrigerant storage for absorption cooling is feasible, but it would be more viable if it is integrated with other storage options for absorption systems. The system can be optimized for reduced solar collector areas and initial solution mass required for continuous operation. The auxiliary refrigerant mass can be evaluated to determine the size of the storge tanks and/or alternatively the fraction of the cooling load being met by solar energy. It is estimated that
206
KAUSHIK et aL: AQUA-AMMONIA ABSORPTION COOLING SYSTEM Table 2. Mi = M^si + Mcsi. Hourly mass stored and mass required/available for matching the cooling load Time (h) 8.00 9.00 10.00 11.00 12.00 13.00 14.00 15.00 16.00 17.00 18.00 19.00 20.00 21.00 22.00 23.00 1.00 2.00 3.00t 4.00 5.00 6.00t 7.00
MAS 827.88 783.20 727.57 672.65 625.83 595.83 577.00 588.85 622.62 654.64 684.46 711.72 736.20 757.85 777.00 793.76 823.00 836.70 850.77# 865.52 882.00 900.53t 921.22
Mcs 22.12 66.82 122.43 177.34 224.16 254.16 273.20 261.14 227.38 195.35 165.53 138.28 113.81 92.15 73.10 56.24 27.10 13.33 49.22# 34.50 18.10 49.50t 28.80#
MASREO MASAVAL 21.76 24.61 27.50 30.50 32.77 34.08 34.54 33.94 33.77 32.02 29.82 27.26 24.46 21.67 19.07 16.84 15.10 14.08 14. I 1 14.74 16.41 18.61 20.70
43.88 69.31 83.11 85.42 79.59 64.08 53.45 22.01 00.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00
tDenotes the case when Mcs = 0, some auxiliary refrigerant mass (50 kg) is added in the condenser store to sustain the system operation. a r a t i o o f the initial s o l u t i o n m a s s to the s o l a r c o l l e c t o r a r e a o f the o r d e r 6 - 8 k g / m : a p p e a r s to m e e t the c o o l i n g l o a d o n a d a i l y cycle.
Acknowledgements--The authors gratefully acknowledge the financial support from T.E.R.I. (India). One of us (SCK) also wishes to thank Dr N. R. Sheridan for his initial association on the dynamic simulation of heat pump problem and Prof. M. S. Sodha for his overall guidance.
REFERENCES 1. N. R. Sheridan, On solar operation of absorption air-conditioners. Ph.D. thesis, Univ. Queensland, Brisbane, Australia (1968). 2. P. C. Auh, A survey of absorption cooling technology in solar applications. Department of Applied Science, B.N.L.-50704, Upton, New York (1977). 3. J. A. Duffle and W. A. Beckman, Solar Engineering Thermal Process. Wiley, New York (1980). 4. J. Baughn and A. Jackman, Solar energy storage within the absorption cycle. ASME publication paper-7H-WA/HT-18 (1974). 5. P. J. Wilbar and C. G. Mitchell, Sol. Energy 17, 193 (1975). 6. M. H. Semmens, P. J. Wilbur and W. S. Duff, Liquid refrigerant storage in absorption air conditioners. ASME-74 WA/HT-19, ASME, New York (1974). 7. S. L. Grassie and N. R. Sheridan, Sol. Energy 19, 691 (1977). 8. W. Bessler and C. N. Shen, Study on parameter variations for solar powered lithium-bromide absorption cooling. Records of the lOth Intersociety Energy Conversion Engineering Conference, University of Delaware, New York (1975). 9. E. Brousse, B. Craudel and J. R. Martine, Appl. Energy 14, 131 (1983). 10. M. O. McLinden and S. A. Klein, Sol. Energy 31, 473. 11. S. C. Kaushik, N. R. Sheridan, K. T. Lam and S. Kaul, J. Heat Recovery Systems, 5, 101 (1985). 12. S. C. Kaushik, Solar absorption refrigeration and space conditioning: thermal modelling of aqua-ammonia cycle. In Reviews of Renewable Energy Resources. Vol. 1, Chap. 4, pp. 230-323. Wiley Eastern, India (1982).
0196-8904/91 $3.00+ 0.00 Copyright © 1991 Pergamon Press pie
Printed in Great Britain. All rights reserved
D Y N A M I C SIMULATION OF AN A Q U A - A M M O N I A ABSORPTION COOLING SYSTEM WITH REFRIGERANT STORAGE S. C. KAUSHIK, S. K. RAO and RAJESH KUMARI Centre of Energy Studies, Indian Institute of Technology, Hauz Khas, New Delhi-110 016, India
(Received 18 September 1989; receivedfor publication 30 August 1990) A b s t r a c t - - A dynamic computer simulation model of a solar powered a q u a - a m m o n i a absorption cooling system with refrigerant storage is presented. The model requires a known solution flow-rate as well as average hourly weather data and the cooling load pattern as input parameters and is capable of predicting various heat flows, temperatures, concentrations, mass flows and mass stored in the absorber/condenser stores. The design parameters of the system components are evaluated at rated conditions using the steady state model. Test simulation of the dynamic model is carried out with the cooling load corresponding to a typical cold storage house on a hot summer day. It is found that the refrigerant stored during excess generation hours m a y not be sufficient to meet the cooling load during off-generation hours, and hence, provision to evaluate the auxiliary refrigerant mass required and the fraction of the load met by solar energy is included in the simulation model. Thus, the concept of refrigerant storage within the cycle is feasible, but it has to be integrated with other storage options for absorption cooling systems. Absorption cooling Dynamic simulation ous space conditioning
Aqua-ammonia
Refrigerant storage
NOMENCLATURE A = Cp = E = h = ~r = MPR = P = Q = t = U = u = v= x =
Area (m 2) Specific heat of water/air (kJ/kg°C) Internal energy (kJ/kg) Enthalpy of fluid (k J) Mass flow-rate of refrigerant/absorbent (kg/s) Mass pumping ratio Pressure (kPa) Heat transfer rate (W/m:) Temperature (°C) Heat transfer coet~cient (kW/m 2 °C) Specific internal energy (kJ/kg) Specific volume (m3/kg) Concentration of refrigerant in solution (kg/kg)
Subscripts a = A= AS = C = CS = E = G = i= 1= o= p= r= R = S= s= V= W = w= I, II = 1-17 =
Air (ambient/cooling space for low temperature heat supply) Absorber Absorber store Condenser Condenser store Evaporator Generator Inlet collector fluid Liquid phase Outlet collector fluid Collector parameter Refrigerant Rectifier Solar radiation Strong solution (rich in refrigerant) Vapour phase Water Weak solution (poor in refrigerant) Type of heat exchanger (preheater/subeooler) State points
Greek letters 0 = Time (h) = Absorptivity ECM 32/3---A
197
Continu-
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z = Transmittance E = Nonequilibrium approach factor r / = Effectiveness of heat exchangers
INTRODUCTION In solar operation of absorption space conditioning [1-3] systems, heating/cooling is accomplished by solar energy during sunshine hours and by an auxiliary source of energy during off-sunshine hours. Since the supply of solar energy and the demand of heating/cooling load are time dependent with certain phase lags, energy storage becomes essential for the efficient use of these absorption cycle systems. Several types of energy storage, viz. hot water storage or cold water storage etc., are possible in conjunction with solar operated absorption systems, but these storage systems have low storage capacity per unit volume in addition to higher thermal loss to ambient and are bulky in nature, and hence, these storage systems can be used for domestic size solar space conditioning. The use of phase change storage media can reduce the storage volume, but their performance gets degraded by 'repeated usage. Recently, a novel concept of solar energy storage by refrigerant storage (utilizing the latent heat of evaporation within the absorption cycle itself) has been proposed by a number of workers in the literature [4-6]. Thereafter, several authors [7-10] have studied the refrigerant storage concept for solar cooling. The working fluid pair in their study was LiBr-H20. However, it is also worth studying the NH3-H20 absorption cooling system, since it has some potential advantages (like no-crystallization etc.) over the LiBr-H20 system [11, 12]. In the present communication, the authors have carried out a dynamic simulation of an NH3-H~O absorption cooling system with refrigerant storage to study the time dependent system performance for matching of the cooling load corresponding to a typical hot summer day. AN
ABSORPTION
AIR
CONDITIONER
WITH
REFRIGERANT
STORAGE
The system considered is shown in Fig. 1. It includes the generator, rectifier, condenser, evaporator and absorber along with preheating/subcooling heat exchangers, pump and pressure ] I "1"C2 I
I-
I
~ /
Condenser
~- I
I I t 12
TRSZ ~ l _ ,, C..T. store~"~'~cs _ i
Refrlgeront
13
r~s~
exchonoer
19
t,i l!
1
Fig. 1. An absorption airconditioner with refrigerant storage.
I I I I I J
KAUSHIK et al.: AQUA-AMMONIAABSORPTION COOLING SYSTEM
199
reducing values. In refrigerant storage absorption cycle systems, a storage volume is provided in association with the condenser to store the refrigerant during the hours of excess generation (i.e. hours of high insolation). The stored refrigerant is released in a manner to meet the required load during off-generation hours. Storage is also needed in the absorber to accommodate not only the refrigerant but also sufficient absorbent to keep the concentration within allowable limits. The condenser store is assumed to have a regulating value at its outlet to control the refrigerant flow. Heat input to the generator is supplied through a flat plate solar collector. The heat of condensation at the condenser and the heat of solution at the absorber are rejected to the atmosphere by a cooling tower, from which water circulates through the absorber store, absorber condenser store and condenser. Room air is circulated through the evaporator at constant inlet temperature. SYSTEM MODEL AND BASIC EQUATIONS In addition to the usual assumptions [11] made in modelling any solar powered absorption cycle system, the following are the two extra assumptions made for modelling the solar powered absorption cooling system with refrigerant and absorber stores. (i) The storage rate remains constant over the time interval under consideration, and the total storage for the interval can be found by first order approximation, (ii) The mass of refrigerant at any time in components other than the stores is neglible. Modelling of different components of the system is discussed next. Solar collector
A fiat plate collector supplies the heat input to the generator. The flat plate collector is modelled according to the collector performance equation [12] given by Qp = Fp {(ez)Q~(0) - Up[ti - ta(O)lAp.
(1)
This Qp is also equal to the heat taken by the circulating heat transfer fluid Qp = 3;/pCw (to - ti).
(2)
The useful heat from the collector is transfered to the generator through the heat exchanger relation. Therefore, we have
Qp=QG=(UA)c
[(to -- tc) -- (ti -- tc)] lnr(to_ tc) 1
(3)
/(ti- tc)_] Generator
The generator is assumed to be a heat exchanger with the collector heat transfer fluid flowing on one side of the heat exchanger and a two-phase flow on the other side. Since a larger part of the generator heat is utilized for boiling of the generator solution, the lower side of the heat exchanger is assumed to be at outlet temperature. The residential time in the generator is too small for equilibrium conditions to be set up. Therefore, an equilibrium approaching factor E is defined as
Xs-Xw E
Xs- x*
(a)
the ratio of the differences in concentrations of the ammonia solutions leaving the generator and absorber. E = 1 corresponds to the full equilibrium state. Here X* is the equilibrium concentration of the solution in the generator at the pressure and temperature of the generator. The basic energy balance equation for the generator is {~G "JI- 2~f,h 5 -3t- .~rl0hl0 = J~rrh 6 --[- 2~/9h 9
where the state points correspond to Fig. 1.
(5)
200
KAUSHIK et
al.:
AQUA-AMMONIA ABSORPTION COOLING SYSTEM
Rectifier The rectifier is also a heat exchanger. It is assumed that the temperature throughout the rectifier is the same. The following are the energy balances around the rectifier
-~fgh9 =/~fl0hl0 q" ~fll hll + QR
(6a)
QR = A:/'4h, - A:/'5h5 = (UA)R [(tz -- ts) -- (tR -- t,)] l n ( t R - ts'~ \tR - t4}
(6b)
In the above equations, QR is the heat evolved at the rectifier. This heat is normally fed to the strong solution before it enters the generator to reduce the generator heat input requirement. Condenser The condenser is also modelled as a heat exchanger. However, the process in condenser heat transfer is different from that in the other heat exchangers of the system because of the condensation of the refrigerant. The UA product of the condenser heat exchanger becomes a function of the refrigerant flow-rate as the heat transfer due to condensation varies over a wide range with flow-rate and, hence, with the thickness of the condensate. Its UA product varies with the flow-rate of the refrigerant since the thickness of the condensate and, hence, the heat transfer rate vary with flow-rate. An equation developed by Grassie and Sheridan [7] is used to determine the actual UA product of the condenser 3.77 (UA)c = 2.4 + 5.5 MIt 3 (UA)r~
(7a)
where (UA)s is the nominal rating of the condenser. The above equation holds good only for small variations over 3.0 kW/s of (UA)N. For the condenser, one can write the following energy balance equations. A;/llh~ = M12h12 + Qc
(7b)
Q¢ = AlaCw(t¢2 - t c l ) = (UA)c [(tc- tel) - (to - to2)] ln(t¢-- tcl~ \t¢ - tc~}
(7c)
Condenser store Mass and energy balances on the condenser store are given by two first order differential equations dMcs = A;/i~ _ A:/l3 dO
dEcs
Ml~hl2 = Might3 + ~
+
(8a)
Qcs
dEcs dMcs . dues dO = dO ucs + M c s dO
(8b)
(8c)
where u is the specific internal energy and is obtained from the state equation (u = h - p v , v is the specific volume) and ....
[(tcs -- tcsl) -- (tcs -- tcs2)]
(8d)
K AU S H I K et al.: A Q U A - A M M O N I A ABSORPTION COOLING SYSTEM
201
The concentration of the refrigerant leaving the condenser store is calculated from the mass and ammonia balance conditions and is given by x5 =
M . 12,zx v . 12 ~ J i .CS - I X'~f 1 T ,trJ
(8e)
M'~2 + M~:s i
Subcooler
The subcooler is a liquid-vapour heat exchanger and is characterized by the effectiveness given by hi6 - hi8 r/n = hi 6 - his,13
(9a)
where hts, l 3 is the enthalpy of the vapour at the concentration of state 18 and temperature of state 13. The energy balance around the subcooler is (9b)
~1f13 hi3 "l-/~fl6hl6 =/~f14h14 -q- 1~flshls. Evaporator
The evaporator is modelled as a heat exchanger with the following assumptions. (i) Purge liquid leaves the evaporator at room air temperature rather than the evaporator temperature, (ii) An expansion valve can be designed to allow the required flow from the condenser store, and the refrigerant is evaporated at constant temperature corresponding to the low side pressure set by the equilibrium conditions in the absorber. The energy balance equations for the evaporator are
QE + Mishls = Ml6hl6 ']- Mi7h17
(10a)
.... [(tE, -- tE) - (tE2 -- tE)] Oe=J~IECw(tEI--tE2)=IU/I,E ln---f-~-E7 ~- )---tE ~
(10b)
L(tE2-
tE)J
Absorber
Equilibrium conditions are assumed to exist in the absorber heat exchanger. The low side pressure is determined by the strong solution concentration at state 1 and the temperature of the absorber. The basic equations for the absorber are QA + ,~/l hi = ~;18h8 + -~t19h19
QA = M , Ca(tm -- tAI) = (UA)A
[(tA- tA, ) - ( t A - tA2)] In[/A/- tAl/"]
(1 la) (1 lb)
LtA -- tA2_] Absorber store
The absorber store is similar to the condenser store. The mass and energy balances in the absorber store are given by two first order differential equations which can be solved by first order approximation, dMAs = ,~t t - M 2 dO
(12a)
.~rl hi = -,f/2h2 --t dEAs dO
(12b)
dEAs dMns dO
-
dO
dU,o
UAS + ~
OU
MAs.
(12c)
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KAUSHIK et al.:
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The concentration of the strong solution leaving the absorber store is determined by mass and ammonia balances by assuming that the mass accumulated in the components other than the condenser and absorber stores is negligible, i.e. Mi = MAs + Mcs Xs =
(13a)
MiXi - Mcs X,3
(13b)
nAs
The solution heat exchanger The solution heat exchanger is modelled as a liquid-liquid heat exchanger whose effectiveness is given by h 6 -- h 7
rh = h6 - h7,3
(14a)
where h7,3 is the enthalpy determined at X7 and t3. The energy balance around this heat exchanger can be written as (14b)
M3h3 @/~/'6h6 = J~4h4 d-/l:/Th 7 .
Cooling load A cold storage house cooling load, varying sinusoidally, with a peak variation of 3 kW about a mean load of 7 kW is considered. This load corresponds to the house having area of approx. 250m 2 (living space) and having a heat loss coefficient of UB= 1.5W/m2°C in Brisbane (Australia) [7] and is shown in Fig. 2. The analytical representation of the cooling load considered is QEDES. =
7.0 - 3.0 cos[n (0 - 2)/12] in kilowatts.
(15)
Cooling tower model In the cooling tower model, it is assumed that the outlet water temperature can be expressed empirically as a quadratic function of both the inlet water temperature and ambient wet-bulb temperature. " ,2 . t 2 tw2 = K, + K:'twb + K, + t~b + K,t~, + K,t~, + Krtwb" tw, + K, t2wb't,i + Kstwb t2' _L T k"*''wb w,"
32
lO
3O
\\
28
1.0
3
o. o.e
24
2
22
20
)
I 4
f
I 8
I 12
16
20
Time (h)
Fig. 2. Hourly solar radiation, ambient air temperature and cooling load.
0.2
24
~
KAUSHIK et al.: AQUA-AMMONIA ABSORPTION COOLING SYSTEM
203
The constants K l - g 9 are evaluated by determining the cold water temperature, tw2 for three hot water temperatures, tw,, at each one of three wet bulb temperatures, twb. There are, thus, nine values o f the function tw2(twb -- twt) at independent values o f the variables twb and twt; hence, the K~ - K 9 can be calculated by a simple inversion matrix. NUMERICAL
COMPUTATION
AND S E L E C T I O N
OF P A R A M E T E R S
It is a rather difficult task to model the dynamic operation of an a q u a - a m m o n i a absorption cycle with all its ancillary equipment in a single stage. The calculation procedure for steady-state operation o f a simple absorption cycle, given the c o m p o n e n t temperatures and refrigerant mass flow-rate, is well established and is similar to that for any t h e r m o d y n a m i c cycle. An extension o f this steady-state analysis for the refined absorption cycle with integral refrigerant storage has been m a d e using appropriate assumptions in the present communication. The first step towards modelling the system is to identify the possible input conditions of the cycle. The input variables in the present case are assumed to be the cooling load, solar radiation and ambient conditions along with the refrigerant concentration and the solution mass flow-rate. The effectiveness of the solution heat exchanger and the subcooler and nonequilibrium a p p r o a c h factor in the generator are also assumed to be known. Initial estimates of the parameters can be m a d e on the basis of some assumed steady-state operations; however, the machine will rarely operate at such steady state conditions, During the day, considerably m o r e refrigerant will be generated than is immediately required to meet the cooling load, whereas overnight, the absorber and e v a p o r a t o r must still operate to supply the cooling load, yet no generation will take place. F o r design calculations, fixed cold storage r o o m air temperture ( ~ 10°C) and fixed supply inlet water temperature (-,~ 20°C) were chosen, the (UA) values o f the various heat exchangers were chosen to give a log mean temperature difference o f 6.6°C when transferring its m e a n load to the cooling water/air, and later, all other UA values and mass flows were adjusted to give consistent results. The values of the p a r a m e t e r s chosen in the simulation model are as shown in Table 1.
Collector parameters Ap = 9 0 m 2, (~r) = 0.9, Fp = 0.9, Ue = 4.4 W / m 2 °C, ~t;/p = A;/Gw.
Building parameters AB = 250 m 2, UB = 1.5 W / m 2 °C.
Cooling tower constants (following Grassie [7]) Kt = -0.21285547E3,
/(2 = 0.50199585E1,
K3 = 0.22859833E - 1,
K4 = 0.72521973E1,
K 5 = - 0 . 4 3 2 7 5 3 5 6 E - 1,
K6 = -0.15433693,
/(7 = 0.92089176E - 3,
/(8 = 0.10446012E - 2,
K9 = 0.64719352E - 5.
Auxiliary mass needed = 100 kg (after midnight 2.00 h). DISCUSSION
OF RESULTS
T o have a numerical appreciation of the concept of refrigerant storage for solar air conditioning, a steady-state simulation based on mass and energy balance equations of different c o m p o n e n t s of Table 1. Selection of parameters Parameter value Parameter value Parameter value (UA)G = 5.625 kW/°C E = 0.8 ~/Gw = 0.95 kg/s (UA)AB= 3.8 kW/°C rh = 0.8 M.A~ = 1.115 kg/s (UA)c = 3.8 kW/°C ~/n= 0.7 Mew = 1.115 kg/s (UA)R = 1.4 kW/°C MABi = 850 kg MEA= 1.0 kg/s (UA)e = 0.75 kW/°C Xi = 0.5 h':/Asw = 1.115 kg/s (UA)AS= 1.2 kW/°C 5;/s = 0.0015 kg/s m2 ~tcs w = I. 115 kg/s (UA)cs = 0.3 kW/°C Cpw= 4.198 kJ/kg°C CpA= 1.005 kJ/kg°C
204
K A U S H I K et al.:
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the system (excluding stores) has been developed to predict the approximate values of the design parameters. The thermodynamic properties of aqua-ammonia are used in the form of state equations expressed by earlier authors [12]. The input parameters for this model include input heat supply, constant room/ambient air temperatures, mass flow-rate of strong solutions, UA values of the component heat exchangers, effectiveness of preheater and subcooler heat exchangers, mass flow-rates of heat transfer fluids, non-equilibrium approach factor in the generator (generator is assumed to be under nonequilibrium state for the given mass-flow of the strong solution) and the design parameters of the flat plate collector. Based on the steady-state analysis, the model has then been extended for dynamic simulation of an absorption air conditioner (following Kaushik et al. [11]) including the refrigerant and absorbent stores within the cycle. To study the feasibility of refrigerant storage within the aqua-ammonia absorption cycle, a test simulation is carried out for a typical hot summer day in Brisbane (Australia). The typical hourly data for the same day for solar radiation, wet and dry bulb ambient air temperatures and the cooling load of a building are shown in Fig. 2. The dynamic simulation model uses hourly solar radiation, ambient (wet/dry) air temperatures and the desired cooling load as the input hourly variables and predicts the system component temperatures, various heat flow rates, mass flow-rates, solution concentrations and the mass stored in the condenser/absorber. The matching of the cooling load with the available cooling produced is also evaluated to determine the auxiliary refrigerant mass needed to operate the system during the off-generation period by careful control of the refrigerant flow from the store/auxiliary refrigerant mass source. Figures 3 and 4 show the hourly variation of the system component temperatures and heat transfer rates, respectively, for matching the typical cooling load corresponding to a hot summer day for which the hourly input data have been used in the simulation model. The following observations can be made from these figures: (i) For a maximum dry bulb temperature of 33°C and wet bulb temperatures of 25°C, the cooling temperature of 10°C can be achieved. The evaporator temperature is a minimum when the availability of solar radiation and dry and wet bulb temperatures are maximum. (ii) The maximum generator temperature is about 130°C, while the condenser temperature is about 42°C and the absorber temperature is about 34°C.
-60
160
'°t
50
140
4o
n
TG,To :~
120
30
--~TAB
l
,o! 80
T
60
0
I
I
I
I
I
I
I
3
6
9
12
15
le
21
Time (h) Fig. 3. Hourly variation o f component temperatures.
:0 24
K A U S H I K et al.:
A Q U A - A M M O N I A A B S O R P T I O N C O O L I N G SYSTEM
205
30
50
25 4.0
ca
E QA8
QG
Qc
15 QE
30
~. p
lo
Q) -r
•
"1"
20
10
O
3
6
9
12
I5
18
Zl
24
Time (h) Fig. 4. Hourly variation of component heat transfer rates.
(iii) The generator heat input, rectifier heat and condenser heat rejection are finite during sunshine hours, while the absorber heat and the evaporator cooling load are finite for all hours. It is also seen that QABis a maximum around 17.00 h, while QE is desired to be a maximum around 14.00 h. (iv) The generator and rectifier temperatures during the non-generation period are just the extra-plotted dashed curves, as there is no justification to evaluate these during these hours. It must be mentioned here that, in order to maintain a constant building temperature, the cooling available from the absorption air conditioner should match the building cooling load. When the cooling available is more than the load, the excess refrigerant generated is stored in the condenser store, i.e. the cooling rate is reduced to that fraction of the time step which would make the cooling produced equal to the cooling required for the total time step. For this portion of the time step, the machine used whatever solar energy was available to generate and store refrigerant. When the cooling rate becomes less than the load, the stored refrigerant is released to meet the load. If the refrigerant available, both from the store and from what is then being generated, is insufficient to sustain the load, the cooling rate is controlled by the maximum refrigerant flow-rate available from the store, and also auxiliary energy could be used to meet the load. However, when the refrigerant mass stored in the condenser store is reduced to zero, then auxiliary mass of refrigerant is added in the condenser store to maintain the system operation. The hourly variations of the mass stored in the absorber and condenser stores as a function of time are shown in Table 2. It is also possible to evaluate the fraction of the cooling load met by solar power and the auxiliary mass of refrigerant required for continuous operation of the system. It can be seen from this table that, from 8.00 to 2.00 h, the total mass (MAS+ Mcs) is equal to mi = 850 kg. However when Mcs is reduced to zero, it is assumed that the extra refrigerant mass of 50 kg is added to maintain the system operation. This table also shows the hourly variation of the refrigerant mass required to match the load and the refrigerant mass available during the system operation. Thus, the size of the storage tank can be determined from these considerations. Thus, the concept of refrigerant storage for absorption cooling is feasible, but it would be more viable if it is integrated with other storage options for absorption systems. The system can be optimized for reduced solar collector areas and initial solution mass required for continuous operation. The auxiliary refrigerant mass can be evaluated to determine the size of the storge tanks and/or alternatively the fraction of the cooling load being met by solar energy. It is estimated that
206
KAUSHIK et aL: AQUA-AMMONIA ABSORPTION COOLING SYSTEM Table 2. Mi = M^si + Mcsi. Hourly mass stored and mass required/available for matching the cooling load Time (h) 8.00 9.00 10.00 11.00 12.00 13.00 14.00 15.00 16.00 17.00 18.00 19.00 20.00 21.00 22.00 23.00 1.00 2.00 3.00t 4.00 5.00 6.00t 7.00
MAS 827.88 783.20 727.57 672.65 625.83 595.83 577.00 588.85 622.62 654.64 684.46 711.72 736.20 757.85 777.00 793.76 823.00 836.70 850.77# 865.52 882.00 900.53t 921.22
Mcs 22.12 66.82 122.43 177.34 224.16 254.16 273.20 261.14 227.38 195.35 165.53 138.28 113.81 92.15 73.10 56.24 27.10 13.33 49.22# 34.50 18.10 49.50t 28.80#
MASREO MASAVAL 21.76 24.61 27.50 30.50 32.77 34.08 34.54 33.94 33.77 32.02 29.82 27.26 24.46 21.67 19.07 16.84 15.10 14.08 14. I 1 14.74 16.41 18.61 20.70
43.88 69.31 83.11 85.42 79.59 64.08 53.45 22.01 00.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00
tDenotes the case when Mcs = 0, some auxiliary refrigerant mass (50 kg) is added in the condenser store to sustain the system operation. a r a t i o o f the initial s o l u t i o n m a s s to the s o l a r c o l l e c t o r a r e a o f the o r d e r 6 - 8 k g / m : a p p e a r s to m e e t the c o o l i n g l o a d o n a d a i l y cycle.
Acknowledgements--The authors gratefully acknowledge the financial support from T.E.R.I. (India). One of us (SCK) also wishes to thank Dr N. R. Sheridan for his initial association on the dynamic simulation of heat pump problem and Prof. M. S. Sodha for his overall guidance.
REFERENCES 1. N. R. Sheridan, On solar operation of absorption air-conditioners. Ph.D. thesis, Univ. Queensland, Brisbane, Australia (1968). 2. P. C. Auh, A survey of absorption cooling technology in solar applications. Department of Applied Science, B.N.L.-50704, Upton, New York (1977). 3. J. A. Duffle and W. A. Beckman, Solar Engineering Thermal Process. Wiley, New York (1980). 4. J. Baughn and A. Jackman, Solar energy storage within the absorption cycle. ASME publication paper-7H-WA/HT-18 (1974). 5. P. J. Wilbar and C. G. Mitchell, Sol. Energy 17, 193 (1975). 6. M. H. Semmens, P. J. Wilbur and W. S. Duff, Liquid refrigerant storage in absorption air conditioners. ASME-74 WA/HT-19, ASME, New York (1974). 7. S. L. Grassie and N. R. Sheridan, Sol. Energy 19, 691 (1977). 8. W. Bessler and C. N. Shen, Study on parameter variations for solar powered lithium-bromide absorption cooling. Records of the lOth Intersociety Energy Conversion Engineering Conference, University of Delaware, New York (1975). 9. E. Brousse, B. Craudel and J. R. Martine, Appl. Energy 14, 131 (1983). 10. M. O. McLinden and S. A. Klein, Sol. Energy 31, 473. 11. S. C. Kaushik, N. R. Sheridan, K. T. Lam and S. Kaul, J. Heat Recovery Systems, 5, 101 (1985). 12. S. C. Kaushik, Solar absorption refrigeration and space conditioning: thermal modelling of aqua-ammonia cycle. In Reviews of Renewable Energy Resources. Vol. 1, Chap. 4, pp. 230-323. Wiley Eastern, India (1982).