Dynamic simulation of an ammonia-water absorption cycle solar heat pump with integral refrigerant storage

Heat Recovery Systems Vol. 5, No. 2, pp. 101-116, 1985

0198-7593/85 $3.00+ .00 Pergamon Press Ltd

Printed in Great Britain.

D Y N A M I C SIMULATION OF AN A M M O N I A - W A T E R ABSORPTION CYCLE SOLAR HEAT PUMP WITH I N T E G R A L R E F R I G E R A N T STORAGE S. KAUSHIK, N . SHERIDAN*, K . LAM* a n d S. KAUL Centre of Energy Studies, Indian Institute of Technology, New Delhi, India 110016

Almtruct--A detailed computer simulation model of an ammonia-water absorption cycle solar heat pump with integral refrigerant storage is reported. Incident solar radiation, ambient and room temperatures and the constant flow rate of the pump are the input to the computer model based on the usual heat and mass balances. First the system is optimised by running the model at average conditions and selecting the best values of the design parameters--heat exchanger UA-values, mass flows, etc. System component temperatures, energy flows and coefficient of performance can be predicted. Next the dynamic model is developed with hourly average data for a typical house in Melbourne to predict the feasibility of the refrigerant storage for matching the heating load of the house during the off-generation period by careful control of the refrigerant flow from the store. The system is shown to be technically feasible and to be marginally more effective than other storage alternatives. Its economic feasibility will depend on the cost benefits that can be attributed to the advantageous storage volume and the low storage losses.

NOMENCLATURE A C

E F h m

rh q

Q s(o) t

U U X

y £

0 Subscripts a A, AS b B C, CS E G i 1 o P R r S, SS V W, WS

collector area [mrj specific heat [kJ kg-i °C-i] internal energy [1o-] heat removal factor enthalpy of the fluid [kJ kg-l] mass of the fluid [kg] mass flow rate of the fluid per unit collector area [kg s -~ m-r] mass flow rate of the fluid [kg s -I] heat transfer rate per unit collector area [kW m-r] heat transfer rate [kW] solar flux density [kW m-r] temperature [°C] overall heat loss coefficient [kW m - ' °C -~] specific internal energy [kJ kg-i] specific volume [mj kg-i] refrigerant concentration in the solution [kg kg-)] solar absorptance of the collector plate transmittance of the glass over effectiveness of the solution heat exchanger nonequilibrium approach factor time [h] room air absorber/absorber store ambient air building condenser/condenser store evaporator generator initial liquid phase outlet average collector property rectifier refrigerant store and strong solution vapour phase weak solution and store

*Mechanical Engineering, Queensland University, Brisbane, Australia. lOl

102

S. KAUSHIK el al.

Superscript n number of time hours

INTRODUCTION An environmental control system utilising solar energy would generally be more cost effective if it were used to provide both heating and cooling requirements in the building it serves. One of the most promising possible alternatives for both heating and cooling is the use of an absorption cycle based solar system. Although various solar powered absorption airconditioning systems [1-3] have been tested extensively but these are not yet cost effective. Solar powered absorption cycle heat pump systems for heating applications have received very little attention during the past several years because of their low coefficients of performance as compared to the vapour compression cycle based heat pump systems. However, this can be compensated by the fact that these systems can be operated with low grade heat available from solar collectors. The other relative merits/demerits of an absorption cycle solar heat pump over a direct solar space heating system and solar assisted heat pumps have been enumerated in the literature elsewhere [4, 5]. Although these solar energy systems have been shown to be technically feasible for heating/cooling of buildings and a variety of solar equipment are available; however, a cost effective equipment for preferable design is yet to be determined. In solar operation of these absorption cycle systems heating/cooling is accomplished by solar energy available during the day hours and auxiliary source is needed during night hours or during the cloudy day. Since both the solar energy input and the demand for heating or cooling load are time dependent in different fashions, energy storage is an essential component for the efficient operation of these systems. Energy storage is also desirable in absorption cycle solar systems for sizing the heat operated devices more in relation to average load than the peak and providing the reserve for peak loadings. Conventionally solar energy is stored as sensible heat in water or rock in conjunction with solar operated absorption systems. Although such storage systems are practicable for domestic size solar space heating/cooling applications; however, these storage systems have low storage capacity per unit volume, high thermal loss to the ambient and are bulky in nature. Phase change storage media (e.g. salt hydrates and paraffins) with constant storage temperature have also been proposed to provide a means of reducing storage volume (at least by a factor of 20-25) but there are practical difficulties in their use due to their performance degradation on cyclic operations of the storage system. Two possible locations for thermal energy storage can be used in an absorption cycle based solar heat pump (Fig. l(a)) either on input side as high temperature water storage between the solar collector and the generator of the absorption cycle or on output side as low temperature hot air storage between the condensor and/or absorber and the space to be heated. On the input side the possible temperature change of 20°C produces a storage capacity of I0 MJ m -3 for water and thermal losses from such storage are substantial due to the temperature difference of 50-60°C from mean storage temperature to the ambient temperature. A computer modelling study [6] of a space heating system indicated that as the collector and/or the storage size is increased relative to other system components, each new increment of thermal capacity is less effective than previous ones in improving the solar fraction achieved by the system. On the output side in case of low temperature hot air storage an effective temperature change of 10°C would limit the storage capacity to 40 MJ m -3. However, the thermal losses could be low since the temperature difference of 15-20°C between the average storage temperature and the ambient is less than half of that for storage on the input side. In addition, since the converted energy is being stored, less storage volume is required for a given heating capacity than for high temperature water storage. However, with the output side storage only, the heat pump system must be sized to the peak energy input resulting in low utilisation of installed plant capacity and low conversion efficiency at part load operation of the system. Recently a novel concept of solar energy storage by refrigerant storage utilising latent heat of evaporation within the absorption cycle itself has been proposed by a number of workers [7-9]. Basically, the concept of refrigerant storage is to provide in association with the condenser, a storage volume where refrigerant can be accumulated during the hours of high solar insolation. The stored liquid refrigerant is released by expansion into the evaporator at other times

103

Ammonia-water absorption cycle solar heat pump Ambienl air

T EI Absorption : Heat ", : Pump : .~

Hot water storage

QG- =.w+th : '

Hot air storage or water storage

mm , storage '

"1

11 Rectifier

I

)co__r 'oomar ,I

t,

I

'

Refrigerant store

> I '; I " !

Tc~2_ :

Room

air

Tc.~t

r13

Heat exchanger 2

18

4 114

Heat exchanger

~V2 r15 T= 2

l

T'o__L

-

0

Evap°rat~_4_.~

Tab

Room air

Room air

t17

Absorber

TA1 c__ .

.

.

.

.

.

.

.

.

.

.

.

T" _1

Fig. l(a). An absorption cycle solar heat pump with thermal storage as space heating system. (b) An absorption cycle solar heat pump with refrigerant storage.

as necessary to satisfy the required loads. Storage is also needed in the absorber to accommodate not only the refrigerant but sufficient absorbent to keep the concentration within allowable limits. Solar air conditioning with refrigerant storage has been investigated in detail and shown to be technically feasible. However, solar space heating with energy storage has not been evaluated so far. Since the solar space heating systems are more cost effective and heating demand is out of phase with the avilability of solar energy as compared to solar space cooling systems; it is more desirable and worthwhile to investigate the absorption cycle heat pump with solar energy storage for space heating applications. In this communication the authors have investigated the modelling of aqua-ammonia absorption cycle solar heat pump with refrigerant storage for efficient space heating of buildings. The proposed system consists of a solar heated generator, rectifier, condenser, evaporator, absorber, solution heat exchanger (i.e. pre-heater), pre-cooler and storage containers. The heat rejected by the absorber and condenser is used to warm the internal air of the room while the evaporator is supplied with the heat from the environment. The heat rejected by the rectifier is used to heat the strong solution before entering the generator and this reduces the generator heat input requirement of the system. In the storage mode refrigerant is boiled from the solution in the generator solar energy in the day time and condensed for storage in the condenser store. When space heating is needed, the condensed refrigerant will be evaporated in an outdoor air coil before being dissolved back into the solution in the absorber. In the heat pump operation

104

S. KAt.JSriXKet al.

the evaporator absorbs heat at low temperature from the ambient air and heat rejected by the absorber and condenser is used to warm the interior air of the space to be heated. The heat of solution and heat of vaporisation/condensation are the energy available for space heating. A steady state simulation model based on mass and energy balances around the components in an absorption cycle heat pump has been developed to decide the design UA- values of the various components using heat exchanger models and circuit variations such as series or parallel air flow combinations between the absorber and the condenser (storage is excluded in this model). Thermodynamic properties of aqua-ammonia mixture are used in form of state equations as expressed in the earlier investigation [5]. The generator is assumed to be under nonequilibrium state for the given mass flow of the strong solution. The input parameters to the simulation model are the ambient and room air temperatures, constant mass flow rate of the strong solution, UA -values of the component heat exchangers and mass flow rates of the heat transfer fluids, effectiveness of the preheater and subcooler heat-exchangers, the nonequilibrium approach factor in the generator and the generator heat input. Based on the steady state analysis, the simulation model has been extended for the dynamic operation of the absorption solar heat-pump incorporating the refrigerant/absorbent storage with the condenser/absorber. Both the storage and the heat pump modes of operation have been included in the continuous operation of the full cycle to match the heating load of a building in Melbourne for a typical cold winter day. The hourly ambient room air temperature, solar radiation and heating load are the required input variables to the dynamic model. System componentstemperatures, energy flows mass flows, mass storage and the coefficients of performance are predicted in terms of design parameters. Careful attention is given to the control of the refrigerant flow from the condenser store for matching the heat available from the system with the heating load of the building. Typical performance results for a cold winter day in Melbourne are given and finally several recommendations for future investigations are highlighted.

ABSORPTION CYCLE HEAT PUMP WITH R E F R I G E R A N T STORAGE The basic principle of an absorption cycle heat pump is similar to the absorption cooling cycle except that the role of the heated air and cooled air is reversed. In a conventional absorption cycle solar heat pump (which includes solar heated generator, condenser, evaporator and absorber together with a heat exchanger, mechnical pump and expansion valve), strong solution (rich in refrigerant) is pumped from the absorber into the generator at a desired solution concentration. The generator converts strong solution to the refrigerant vapour and weak solution (poor in refrigerant) by the addition of heat from a solar collector. The temperature of the weak solution is higher than that of strong solution and increases the boiling required for the generation of the refrigerant vapour. The refrigerant vapour is condensed rejecting heat in the condenser. The liquid refrigerant on expansion continues to the evaporator where heat is absorbed from the ambient air during the evaporation process. The refrigerant vapour coming from the evaporator then continues to the absorber where it joins the weak solution coming from the generator through solution heat exchanger and the expansion valve and a highly exothermic reaction takes place thereby rejecting heat in the absorber. The resulting strong solution then completes the cycle. The heat rejected by the absorber and condenser is utilised to warm the interior air of the space to be heated. The choice of the refrigerant-absorbent pair to be used for an absorption cycle heat pump with refrigerant storage is somewhat different from that for a conventional absorption refrigeration cycle. The refrigerant-absorbent combination should be such for which (i) (ii) (iii) (iv) (v) (vi)

the absorption capacity of the absorbent is large, the specific heats of the absorbent and the refrigerant are low the heat of exothermic reaction is large, vapour pressure of the absorbent at the generating temperature is low, vapour pressure of the refrigerant at the working temperature is low, latent heat of vaporization of the refrigerant is large.

Ammonia-water absorption cycle solar heat pump

105

Since a large volume of storal~e~must be provided at the high :pressure level of the system (i.e. that of pure liquid refrigerant at the condenser temperature) a refrigerant with low vapour pressure is preferable. Further, high latent heat of the refrigerant is desired to reduce the storage volume for the liquid refrigerant. Aqua-ammonia mixture is found to be attractive in terms of low storage volume. However, the high storage pressure will require high pumping powers and high pressure tankages. The sensible heat contained in the hot weak solution is used to heat the strong solution entering the generator through solution heat exchanger (also known as preheater) and thus reduces the amount of energy input to be supplied to the generator. The use of a preheater between the generator and the absorber as well as storage of weak solution at lower temperature may greatly reduce the high pressure requirement. Similarly a sub-cooler used between the condenser and the evaporator increases the amount of energy extracted by the evaporator from the ambient air. Since some heat has to be rejected from the rectifier, a further refinement possible to improve the performance of an absorption cycle is to utilise this heat in raising the temperature of the strong solution before entering the generator. This also in a way reduces the generator heat input requirement of the system. However, since water is volatile and hence refrigerant vapour is mixed with water vapour which degrades the performance of the system. A complex rectification system consisting of a distillation column, analyser and a separator is required to separate the water content from the refrigerant vapour. An absorption cycle solar heat pump with refrigerant storage is shown in Fig. l(b). Solar energy storage may be achieved within the absorption cycle using solar energy required to separate a refrigerant from an absorbent solution. Basically the idea is to provide storage volume in association with the condenser where the refrigerant can be accumulated during the hours of high solar insolation. Storage volumes are also needed in or near the absorber to store not only the weak solution but also the strong solution, each being available at different temperatures. A schematic representation of the storage mode of the absorption cycle is also shown (Fig. 2(a)). The refrigerant vapour from the generator is condensed in the condenser and heat of condensation is rejected in the condenser to the room air. The condensed liquid refrigerant is stored in the store as the refrigerant storage and is subsequently evaporated to release the latent heat. The energy can be recovered by the highly exothermic absorption of the refrigerant in weak solution. Storage of weak and strong solutions are also needed in or near the absorber to store not only the refrigerant but also the absorbent to keep the solution concentration required for the generation of the refrigerant in the generator. Both the weak and strong solutions can be stored in the absorber store in combined storage mode or in separate storage mode as proposed by earlier workers [11, 12]. In the heat pump mode the liquid refrigerant stored in the condenser store is released at low pressure to the evaporator which absorbs energy from the ambient air and the refrigerant vapour is allowed to mix with the weak solution in the absorber. The heat rejected by the absorber is the heat used to warm the interior air of the space to be heated (Fig. 2(b)). SYSTEM MODELLING

AND BASIC SYSTEM EQUATIONS

It is a rather difficult task to model the dynamic operation of an absorption cycle heat pump with all complexities involved and ancillary components in a single stage. In a system as complex and having as many independent variables as that being considered it is necessary to determine which of the variables will have little influence on the performance of the system and which variables might significantly affect the operation. The former can then be considered as constant and selected accordingly where as the latter have to be varied in such a manner that a point of optimum performance is determined. To facilitate variation of these important parameters and reduce the necessity for recompilation of the computer simulation, all variables which might reasonably be expected to affect the cycle performance are assumed to vary while other parameters which may not affect the performance significantly are read as input data in the simulation model. Therefore, first a steady state simulation of an absorption cycle heat pump is carried out in detail (excluding the storage process) to decide the design parameters--heat exchanger--UA-values and mass flows etc and the circuit variations such as series or parallel air flow between the condenser and the absorber. Such an approach for the solar absorption heat pump will enable the prediction of the coefficient of performance in terms of ambient and room temperatures, solar radiation and

106

S. KAUSrlIKet al.

the system design parameters. The model can be then extended for refrigerant storage by assuming that (i) the storage rate remaining constant over the time interval and the total storage for the interval can be found by first order approximation, (ii) the mass of the refrigerant at any time in components other than the stores are negligible. The system components of an ammonia-water absorption cycle for the dynamic model are discussed separately. The solar collector is the logical choice at which to commence an analysis as it receives the energy input from the solar radiations. The solar collector is modelled on the basis of the Hottel and Woertz [13]. The performance equation for flat plate solar collector is given by ap = Fp [(z~) S(O) - U p ( t , - t.(O))] Ap

(1)

7

l'

enr'"or611

l l

r

Condenser store

all

I, 2

store

Fig. 2(a)

''[/

Generator l I

41

it I

Condenser

I

~12

r

I1" Co.aenser '

'

t•

11

I

Absorber s,ore

h14

18~

Absorber /U

I

I I

liI

1-9

I

I=

2 Fig. 2(b)

Ambient l

Ammonia-water absorption cycle solar heat pump

107

i

.I

c]

A Cs

!s I Jl

I Room air I

l Roomair load

{I) Seriescircuit- A-C- L

(11) Seriescircuit-C-A- L

I I

Absorber

+ Absorber store

I Condenser

]

I

Store

1

m

{111)Parallel circuit Fig. 2(c) Fig. 2(a). Storage mode of an absorption cycle solar heat pump, (b) Heat pump mode of an absorption cycle solar heat pump. (c). Heating circuit variations.

which is also equal to the heat taken by the circulating heat transfer fluid Qp = ~,lp C,, (to - t~).

(2)

Since the useful heat available from the collector is transferred to the absorption cycle working fluid in the generator through the heat exchanger relation: [(t o - tG) - ( t , - tG)]

In[~t°~ tG)ltG)_j

QP=QG=(UA)G

(3)

where the generator temperature is assumed to be uniform along the generator side. The g e n e r a t o r is modelled as a heat exchanger with collector heat transport fluid on one side and a boiling two phase flow on the other. When the absorption cycle beat pump is generating refrigerant by far the greater part of the energy is used for vaporising the refrigerant along with a relatively small proportion being used for sensible heating of the solution; for this reason the low side heat exchangcr temperature was assumed to be constant at the outlet temperature. This approximation is used in the absence of more exact information on the temperature variation in the generator. The boiling point of the solution will actually vary due to decreasing concentration of the solution as the refrigerant is boiled from it. Furthermore, it is necessary for the nonstcady state operation to assume an approach to equilibrium in the generator. The weak solution leaving the generator approaches the equilibrium state corresponding to the temperature and pressure of the generator. As the residence time in the generator is insufficiently long for equilibrium conditions to be achieved, an approach to equilibrium factor E is used where E is defined as the ratio of the difference in concentrations of ammonia in the streams leaving the generator and absorber to the maximum possible difference: E =

x,-Xw x,-x*

~

(4)

where X* is the equilibrium concentration in the generator at the pressure and temperature of the generator.

108

S. KAUSl-IIKet

al.

Thus E is a fractional approach to the equilibrium in the generator and can be considered physically as being analogous to the isentropic efficiency of the conventional compressor in the sense that a cycle with E less than unity will generate less refrigerant for a given heat input than a corresponding cycle with E equal to unity. The basic mass, ammonia and energy balance equations for the generator are given as h;/5 + -Jl:/lO= -~¢6+ M9

(5a)

2~/5X5 "{- 2~/ioXlo = -~6X6 + )~/9X9

(5b)

QG +/~/5h5 "[-/~/10 h,o =/~6h6 +/~9h9

(5C)

where state points refer to Fig. l(b). The rectifier is also modelled as a heat exchanger with uniform rectifier temperature and used to separate the water content from the refrigerant vapour which otherwise degrades the performance of the system. Since some heat has to be rejected from the rectifier a further refinement possible to improve the performance of an absorption cycle is to utilise this heat in raising the temperature of the strong solution before entering the generator. Thus the available rectifier heat is used internally to heat the solution and hence reduces the required generator heat input of the system. The usual equations governing mass, ammonia and the energy balance around the rectifier are given by 5;/9 = A;/10+ A;/. (6a)

2~/9,¢~9 = /~/IoXIo "~ 2~/112~'11

(6b)

-~9h9 = 21;/mh~o+ -h:/uhu + QR

(6c)

. . . . {(tR - t5) - (tR - t4)] Q, = A:/,h4- .b:/',h, = ( u A ) , ]-n[(~ ~ t~(-~-R--- t4)l"

(6d)

The condenser is modelled as heat exchanger to supply heat to the room air which is circulated on one side of the heat exchanger. As the refrigerant flow rate and hence the thickness of the layer of the condensate vary over a wide range, the UA product of the condenser becomes a function of that flow rate. An equation developed from Nusselt [14] Relation to describe this variation is 3.77

(UA)c = 2.4 + 5.5 MI~3(UA)u

(7a)

Where (UA)N is the nominal rating of the condenser. The above empirical equation results from a design value of the conductance area product of 3 kW °C- 1 and is valid only for small changes of (UA)u from this value. The basic equations are /I,:/12= A;/.

(7b)

X12 = Xll

(7C)

/~/ll hll ----J~/12hi2 -I--QC

(7d)

Q C = ]~a Ca(to2 - - lcl) = ( U A )c [(to -

tcl) - (to -

inF(t

/c20]

l

(Te)

L (tc- t~2)J Mass and energy balances on the condenser store yield two first order differential equations (which can be solved numerically with a first order approximation) dMc, = A;/I2 dO

_

A:/,3

(8a)

J~,2hl ':' = J~13hl3 + ~~ + Qcs

(8b)

dE~, dM~, due, d---O = dO "Uc'+M~'"'dO

(8c)

Ammonia-water absorption cycle solar heat pump

109

where u is the specific internal energy and obtained from the state equation (u = h - Pv; v is the specific volume) and Oc, = ( UA )~,

[(t~, -- t~,,) -- (t~,- t~2)] ln[(t¢, -- t¢,l)/(t, -- t,~)]

(8d)

The concentration of the refrigerant leaving the condenser store is calculated from the mass and ammonia balance conditions and is given by

[-M,, .

X73 = L

.

.

.

+ M~",--.IX73-I] .

i

M72 + Mc~-

(8e)

where n is the time hour number The sub cooler is a liquid-vapour heat exchanger and is characterised by the effectiveness given by hi6 -- his r/n = hie - h18.13

(9a)

Where hlsa3 is the enthalpy of the vapour at the concentration of state 18 and temperature of stage 13. The energy balance around the sub-cooler is given by A;/13h13 + A]tl~ha6= -~/14hi4 + Alnhls.

(9b)

In the low pressure side of the system, refrigerant vapour from the evaporator after being heated in the sub-cooler passes to the absorber where it is dissolved in the weak solution returning from the generator. The purge liquid coming from the evaporator also goes to the absorber separately. The assumption made is that the purge liquid leaves the evaporator at the ambient air temperature rather than the evaporator temperature; this is justified on the ground that the liquid is closely in contact with the evaporator coil being at the ambient air temperature. Since the evaporator is also modelled as a heat exchanger with ( U A ) e as a constant, although ( U A ) e could be expressed as a function of refrigerant flow if desired. The liquid refrigerant entering the evaporator is assumed to be evaporated at a constant temperature corresponding to the low side pressure set by the equilibrium conditions in the absorber. It is assumed that an expansion device can be designed to allow the required refrigerant flow. If a conventional expansion valve of constant area were used, the refrigerant flow rate, being controlled by the pressure difference across the value, may often be greater than the amount which could be evaporated and hence such a design is unsuitable in this application. The basic equations for the evaporator are K/15 = A:/'16+ Ml7

(10a)

/~f15 X "¢~15= ./~/16X16 "4"I~rlTXI7

(IOb)

Qr + A:/'15hi5 = A':/zrhl6 +/~17hl7

(10c)

Q r = h'lbCa(tE, - tr2) = (UA)r [(tel - t E ) - ( t e 2 - tr)]

(lOd)

F(',,-',)lt,)J

In L(te'

The absorber is envisaged as a heat exchanger on one side of which room air flows entering at constant temperature. On the other side, the weak solution coming from the generator is distributed over the heat exchanger surface and dissolves any refrigerant vapour coming from the evaporator. The room air reduces the absorber temperature (by extracing the heat of mixing and heat of condensation of the refrigerant vapour) and improves the refrigerant solubility in the absorber. As equilibrium conditions are assumed to exist in the absorber, the low side pressure is that corresponding to the strong solution concentration at state 1 and temperature of the absorber. The basic equations for the absorber are 2~fl "~- ~ 8 ''[" /~f19

(lla)

~flXl =/~f8X8 -[-2~fl9Xl9

(IIb)

I I0

S. KAUSHIKet al. QA -F -'~fl hi = / ~ 8 h 8 Jr -~t19h19

(l lc)

QA = A:/aCa(ta2 - t~]) = (UA)A [(tA -- ta,) -- (tA -- tA:)]

(1 ld)

In[((t-~ ~ tA')l

tA:)J

The absorber store is similar to the condenser store and stores not only the refrigerant but also the absorbent sutticient to keep the allowable concentration difference between the weak and strong solution. Mass and energy balances on the absorbent store give first order differential equations of the same form as those developed for the refrigerant store with the exception that there is assumed to be no heat rejection term equivalent to Qc,. dMAs = A;/I _ ,~t2 dO

(12a)

A~/lht = 3;/2h2 "t dEAs dO

(12b)

dEAs dMAs . dUAs dO - dO "UAs+ MAS" el0"

(12C)

These equations are solved using first order approximations. The concentration in the absorber store is calculated by the mass and ammonia balance conditions and by assuming that there is negligible mass accumulation in components other than the absorber/condenser stores. Thus

X, =

(Mi X i - McsXI3)

(12d) (12e)

Mi = MAs + Mcs.

The solution heat exchanger between the generator and absorber is modelled as a conventional liquid-liquid heat exchanger using the NTU-effectiveness approach. The effectiveness '/i for the solution heat exchanger is given by (h6 - hT)

r/ = h6 - h7,3

14

1.4

12

1.2

14 12 ~

,o

(13a)

D

r

y

bulb

1.o

,o

4

Solar radiation

2

2 0

0

4

8

12

16

20

24

Time (hr)

Fig. 3. Hourly variation of solar flux density, ambient air temperature and the heating load of a building for a cold winter day in Melbourne.

Ammonia-water

absorption

cycle solar heat pump

tI!

where h7,3 is at the concentration X7 and temperature of the absorber store h- Also we have A;/3h 3 +

,/~t6 h 6 = ~ f 4 h 4

+

(13b)

2 ~ 7 h 7.

The building heating load has been estimated by approximate methods for a specific building of 100m 2 floor area and overall heat loss coefficient of 1.5 W m -2 °C -1. For a typical cold winter day (July 1) and desired indoor temperature of 2°C for the building in Melbourne (Aust.) the weather data (ambient air temperature), solar radiation and the desired heating load are shown on hourly basis (Fig. 3). SELECTION OF PARAMETERS AND COMPUTER SIMULATION Initial estimates of parameters can be made on the basis of some assumed steady state operation. However, the machine will rarely operate at such steady state conditions, as during the day considerably more refrigerant will be generated than is immediately required to meet the heating load, whereas overnight the absorber and evaporator must still operate to supply the heating load yet no generation will take place. For design calculations room air temperatures of 21 °C and supply air temperature of 35°C were chosen, (UA) of the absorber heat exchanger was chosen to give a log mean temperature difference of 6.6°C when transferring its mean load to the room air and later all other UA-values and mass flows were adjusted to give consistent results. The values of the parameters chosen in the simulation model are shown in the Table 1. The absorber has relatively a higher UA value as necessary to transfer a reasonable heating load in the night time. The generator and condenser heat exchangers were chosen as to reduce the system high pressure and improve the performance when refrigerant generation is taking place. The rectifier heat exchanger is sized to the minimum value as desired by the strong solution flow being heated through it. The mass flow rates of the ambient/roomair are adjusted accordingly. The collector to generator mass flow rate of 0.02 kgm -2 s -j is generally agreed to be close to the optimum value for flat plate solar collectors with reasonable outlet and inlet collector fluid temperatures. The collector area must be sized to give reasonable storage mass. The absorber to generator mass flow of the strong solution is selected so as to give a reasonable mean mass flow ratio ( ~ 20) during the generation. This gives a good idea of the latent heating and the sensible heating of the solution in the generator. Finally the selection of initial concentration and mass of the solution in the absorber is made on the basis of chosen mass flow rate and equilibrium concentration in the generator. A lower limit is placed on the concentration because otherwise it involves large initial mass of the solution and upper limit should be chosen to have reasonable low pressure in the absorber which is also the pressure in the evaporator. In fact the proper selection of M~, X, and the collector area is made by trial basis and then optimum values are chosen in accordance with the reasonable storage mass, low pressure and concentration differential between the absorber and the generator. With the values for collector area, initial concentration and initial solution mass shown in Table 1, it was found that the required load could be met at all times. DISCUSSION OF RESULTS As mentioned earlier first the steady state computer simulation results of an absorption heat pump have been obtained using constant mass flow rate of the solution pump, generator heat, Table I. Selection o f parameters

Parameter

Value

Parameter

(UA)~ = 1.2 k W ° C - I (UA)A= 1 . 4 k W =C-~ (UA)c = 0.9 k W '~C - i (UA)• = 0.3 k W c C - J (UA)E = 0.75 k W ° C - I (UA)As = 0.3 k W ~ C - ' (UA)cs = 0.3 k W ° C - ) As= 1 0 0 m 2 Up=4.4Wm-2

=C-I

Value

~ = 0.8 ~t=0.8 fill = 0.7 M~ = 500 k g X i = 0.5 M s = (0.04-0.08)kg s - ' m - Z F~ = 0.97 Ae=40m 2

Parameter

Value

M.'p = M a~ ,~th = A~tc = C~w = C~ = R o o m Air T~ = (ct~) =

0.8 k g s l.llkgs-t 0.52 k s s 1.11 k g s - I 4.198 k J k g - t ° C - I 1.005 kJ k g - J ° C - ' 21°C 0.9 Ua= 1 . 5 W m -2 o c - I

S, KAUSHIK et al.

112

Table 2. Steady state cycle analysis. (a) Effect of generator heat input. (b) Effort of solution mass flow rate Parallel circuit

Series circuit A-C-L*

Parameter Q~ (kW) 4 8 12 16 (b) Fs (kg s-tin -2 ) 0.025 0.035 0.045 0.075 0,125

C-A-L*

S-F-R*

D-F-R*

CPc

CPAe

Ce

CPc

CPAB

Cp

CPc

CPAB

Ce

CPc

CPAB

Ce

0.36 0.49 0.52 0.54

0.98 0,95 0.91 0.86

1.34 1.44 1.44 1.40

0.40 0.54 0.58 0.60

0.97 0.93 0.87 0.83

1.37 1.47 1.45 1.43

0,39 0.53 0.57 0.59

0.97 0.93 0.88 0.82

1.36 1.46 1.45 1.40

-0.51 0.55 0.56

-0.94 0.89 0.82

1.45 1.43 1.38

0.68 0.64 0.61 0.51 0.37

0.83 0.88 0.90 0.94 0.96

1.51 1.53 1.51 1.45 1.33

-0.65 0.57 0.45

-0.86 0.90 0.94

-1.50 1.47 1.39

0.70 0.66 0.63 0.55 0.43

0.79 0,85 0.87 0.91 0.94

1.49 1.51 1.50 1.46 1.37

0.68 0.65 0.61 0.53 0.41

0.83 0.85 0.88 0.92 0.94

1.51 1.50 1.49 1.45 1.36

(a)

*A-C-L Absorber to condenser to load: C-A-L Condenser to absorber to load; S-F-R Same flow rate: D-F-R Different flow rate.

ambient and room temperature being the input variables to the computer model. The system is optimised by running the model at average conditions and selecting the best values of the design parameters---heat exchanger UA -values, external fluid mass flows etc. and the effects of these input variables on the heating co-efficient of the performance is predicted. Tables 2 and 3 show the variations of the over-all and component co-efficients of performance as .a function of generator heat input, mass flow rate, ambient and room air temperatures for different heating circuit variations (viz series or parallel heat flows between condenser, absorber and the load). It can be seen that a condenser to absorber to load series heating circuit is preferable Table 3, (a) Effect of ambient temperature. (b) Effect of room temperature Parallel circuit

Series circuit A-C-L*

Parameter (a) TAue CC) 5 7 9 (b) TR (~C) 19 21 23

C-A-L*

S-F-R*

D-F-R*

CPc

CP~e

Ce

CPc

CPAs

Cp

CPc

CP~B

Ce

CPc

CPAe

Ce

0.49 0.50 0.52

0.92 0.93 0.94

1.42 1.44 1.45

0,55 0,56 0.57

0.89 0.90 0.90

1.44 1.46 1.47

0.53 0.54 0.55

0.90 0.91 0.91

1.43 1.45 1.47

0.52 0.52 0.53

0.90 0.91 0.92

1.42 1.44 1.45

0.52 0.51 0.50

0.94 0.94 0.93

1.46 1.45 1.43

0.57 0,57 0,56

0.91 0.90 0.90

1.48 1.47 1.45

0.56 0.55 --

0.92 0.91 --

1.48 1.46 --

0.54 0.53 0.52

0.92 0.92 0.91

0.54 1.45 1.43

*A-C-L Absorber to condenser to load; C-A-L Condenser to absorber to load; S-F-R Same flow rate; D-F-R Different flow rate. Table 4. Effect of UA-values

Series circuit Parameter

Parallel circuit

A-C-L* CPc CPAe

Cp

CPc

CPAs

Cp

CPc

CPAs

Cp

CPc

CP~a

Cp

0.51 0.51 0.52

0.93 0.94 0.93

1.44 1.45 1.45

0.56 0.57 --

0.90 0.90 --

1.46 1.47 --

0.55 0.55 0.55

0.91 0.91 0.91

1.46 1.46 1.47

0.53 0.53 0.54

0.91 0.92 0.92

1.44 1.45 1.45

0,51 0.51 0.51

0,93 0.94 0.94

1.44 1.45 1.45

0,56 0,57 0.57

0.90 0.90 0.90

1.47 1.47 1.47

0.55 0.55 0.55

0.91 0.91 0.91

1.46 1.46 1.46

0.53 0.53 0.53

0.92 0.92 0.92

[.45 1.45 1.45

0.51 0.51 0.51

0.93 0.94 0.93

1.45 1.45 1.45

0.57 0.57 0.57

0.90 0.90 0.90

1.47 1.47 1.47

0.55 0.55 0.55

0.91 0.91 0.91

1.46 1.46 1.46

0.53 0.53 0.53

0.91 0.92 0.92

1.45 1.45 1.45

0.51 0.51 0.51

0.93 0.94 0.94

1.45 1.45 1.45

0.57 0.57 0.57

0.90 0.90 0.90

1.47 1.47 1.47

0.55 0.55 0.55

0.91 0.91 0,91

1.46 1.46 1.46

0.53 0.53 0.53

0.91 0.92 0,92

1.45 1.45 1.45

C-A-L*

S-F-R*

D-F-R*

UAE (kW C -~) 0.65 0.75 0.85

UAc (kW~C - ~) 0.80 0.90 1.00

UAD (kWOC- ~) 0.2 0.3 0.4

UA AB ( k W C -~) 1.30 1.40 1.50

*A-C-L Absorber to condenser to load; C-A-L Condenser to absorber to load; S-F-R Same flow rate; D-F-R Different flow rate.

A m m o n i a - w a t e r absorption cycle solar heat p u m p

113

over the absorber to condenser to load series circuit. Moreover both the series circuits are marginally better in performance than the parallel circuit cases (viz same or different flow rate of the external fluid). It is further seen that there is an optimum generator heat input and optimum mass flow rate of the strong solution for each case when the over all co-efficient of the performance is maximum. It is also noticed that absorber heating co-efficient of performance is always higher than the condenser heating co-efficient of performance and the former shows opposite behaviour as compared to the latter. The effect of increasing the ambient air temperature is to increase the component and over all COP values while the effect of increasing the effect of room air temperature is to decrease the component and over all COP values as expected intuitatively. The influences of the component heat exchanger UA-values on the over all and component heating co-efficient of performance are shown in Table 4. It is seen that there is little change in COP values with the increase of the component UA values and thus justifying the optimum UA-values selected for simulation model. To study the feasibility of internal refrigerant storage within the aqua-ammonia absorption heat pump a dynamic computer simulation process is carried out for a typical cold winter day (lst July, 1970) in Melbourne using a typical building heating load and climatic data (solar radiation and ambient air temperature) as shown on hourly basis in Fig. 3. The initial mass and concentration of aqua-ammonia solution in the absorber/absorber store, heat exchanger effectiveness, approach to non-equillibrium factor, UA-values, external fluid mass flow rates and the equillibrium solution mass flow rate are used as the input parameters in the programme of the simulation model. Hourly Solar radiation, ambient air temperature and the desired heating load are used as the input hourly variables in the dynamic simulation model. The simulation model is developed with hourly average data for a typical house in Melbourne to predict the system component temperature, component heat transfer rates, solution mass flow rates and eoncentra.tion and the requirement of refrigerant storage for matching the heat load of the house during the off-generation period by careful control of the refrigerant flow from the store. A series heating circuit (viz absorber to condenser to load) is considered for the dynamic simulation process.

50

120

40

110 '

TG

Tc

]

o v

100

o v

90

o

I--

20 -

..~ 80

70

60 0

6

12

18

24

Time (hr)

Fig. 4. System components temperatures as a function o f time for a specified building heating load. H.R.S.

5/2~

114

S. KAUSltlK et al.

16-0l

QG

12.0

"QA + Qc +AQ

Ag 8.0 O

QA

4.0

4

8

12

16

20

24

Time (hr) Fig. 5. Hourlyratesof heat transferin variouscomponentswithtimefor a specifiedbuildingheatingload. Figure 4 shows the hourly variations of the system component temperatures as a function of time for the same cold winter day. It is seen that the evaporator temperature is always lower than ambient air temperature (shown in Fig. 3) and the absorber/condenser temperatures are sufficient to maintain a warm environment for the room air. The maximum generator temperature required is about 108°C at 13.00 h (scale shown on R.H.S.). The generator temperature during off Sunshine hours as shown by dotted line is hypothetical value. The actual generator temperature in this period will approach towards ambient conditions. Figure 5 shows the various components heat transfer rates as a function of time. It is seen that both Q6 and Qc are finite only in the generation period while QE is finite in the off generation period. The absorber heat is finite at each hour and is sufficient to match the heating load in the off generation period while the condenser heat is much in excess of the heating load during the generation period. For the sake of illustration the total available energy from the absorption heat pump and the desired building heating load are shown in the same figure. It can be seen that the available heat from the system is always higher than the desired heating load and the excess heat during day time may be stored. In order to maintain a constant building temperature, the heat available from the heat pump should match the building heating load. When the heat available is more than the load the excess refrigerant generated is stored in the condenser store, i.e. the heating rate is reduced to that fraction of the time step which would make the heating energy equal to the energy required for the total time step. For this portion of time step the machine used whatever Solar energy was available to generate and store refrigerant. When the heating 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 heating rate is controlled by the maximum refrigerant flow rate available from the store and also the auxiliary energy could be used to meet the load. Figure 6 shows the variation of the refrigerant mass stored in the condenser and the absorber as a function of time corresponding to the matching of the heating load with the available heat. It can be seen that the mass stored in the condenser increases with time in generation period and decreases in the off-generation period while mass stored in the absorber store decreases with time in the generation period and increases with time in the off-generation period. Thus there is a mass transfer from the absorber store to the condenser store in the generation period and vice versa in the off-generation period. The average mass stored in the condenser store is about 60 kg (maximum 120 kg and minimum zero). The average mass stored in the absorber store is 440 kg, and hence the sizes of the heat exchanger used for these stores are selected. A material balance of the system components is always maintained during the mass transfer phenomena from one store to the other.

Ammonia-water absorption cycle solar heat pump

115

540

500

M,,

I

M cs

J 120

A

v

= 450 -

80 A

40O 40

-

i

Io

350 0

3

6

9

12

15

18

21

24

Time (hr) Fig. 6. Mass stored in the cond©nscr and absorber stores as a function o f time.

CONCLUSIONS

Steady state and dynamic computer simulation models of an aqua-ammonia absorption heat pump have been developed to obtain optimum design parameters and assess the feasibility of the system with internal solar energy storage within the cycle itself. It is found that a condenser to absorber to load heating series circuit is preferable over parallel circuit or absorber to condenser to load in series circuit variations. For a typical heating circuit, viz absorber to condenser to load in series, the steady state model is extended to the dynamic model considering time depending solar radiation, ambient air temperature and the building heating load on hourly basis. The concept of internal refrigerant storage within the absorption cycle heat pump seems feasible for efficient space heating. Concentration and temperature controls in refrigerant storage could be achieved by de-coupling the generator heat input with solar collector or by releasing the refrigerant from condenser store to evaporator and the absorber as and when desirable for the continuous system operation. An absorption heat pump with refrigerant storage is fundamentally sound and more marginally effective than other storage alternatives. Its economic feasibility will depend upon the costs benefits that can be attributed to advantageous reduced storage volume and low heat losses from the storage [16-18]. Acknowledgements--The author (S.C. Kaushik) gratefully acknowledges the financial support from Tata Energy Research Institute, Delhi, India and Fellowship from Queensland University, Brisbane, Australia.

REFERENCES 1. A. T. Ellington, G. Kunst, R. E. Peck and J. F. Read, The Absorption Cooling Process, I.G.T. Research Bulletin No. 14 (1975). 2. J. A. Duffle and N. R. Sheridan, Lithium bromide-water refrigerators for solar operation, Inst. Engrs. Aus. Trans. Mech. Chem. Part I, 79 (1965). 3. N. R. Sheridan, On Solar Operation of Absorption Air-Conditioners. Ph.D Thesie, University of Queensland, Brisbane (1968). 4. M. Balsubramaniam, G. L. Schrenk, A. Lowi and J. C. Denton, Solar Thermal Absorption Heat Pump breakeven coefficient of performance, ASME publication paper-74--WA/Energy-2 (1974). 5. S. C. Kaushik and N. R. Sheridan, Computer modelling and thermodynamic assessment of an aqua-ammonia absorption cycle heat pump, Build Environ 16, 209-220 (1981). 6. J. A. Duffle and W. A. Beckman, Solar Engineering Thermal Process, New Edn, John Wiley and Sons, New York (1980). 7. J. Baughn and A. Jackman, Solar Energy Storage Within the Absorption Cycle, ASME publication paper-74-WA/HT18 (1974). 8. P. J. Wilbur and T. R. Mancini, A comparison of solar air conditioning systems, Solar Energy Ilk 569 (1976).

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S. KAUSHIKet al.

9. N. R. Sheridan and S. C. Kaushi, A novel latent heat storage for solar space heating systems: refrigerant storage, Appl Energy 9, 165-172 (1981). 10. S. L. Grassie and N. R. Sheridan, Modelling of a solar operated absorption air conditioner system with refrigerant storage, Solar Energy 19, 691 (1977). 11. G. Schrenk and N. Lior, The absorption cycle solar heat pump and the potential role of thermochemical energy storage, Proc. Second Workshop on the Use of Solar Energyfor Cooling of BuUdings, Aug. 4-6, p. 207 (1975). 12. J. W. Baughn and M. J. McDonald, Theoretical modelling of an ammonia-water absorption cycle with solar energy storage, Proc. American Section ISES Annual Meeting Conf, Vol. 1, pp. 7-25 (1977). 13. H. C. Hottel and B. B. Woertz, The performance of flat plate solar heat collectors, Trans. ASME 64, 91 (1942). 14. W. H. McAdams, Heat Transmission, Third Edn, McGraw Hill, New York (1954). 15. P. C. Jain and G. K. Gable, Equilibrium property data equations for aqua-ammonia mixtures, ASHRAE Trans. 77, 149 (1971). 16. S. C. Kaushik, Solar Absorption Refrigeration and Space Conditioning, Chap 4, Vol. 1, Reviews in Renewable Energy Sources, Wiley Eastern Ltd, India (1982). 17. M. O. McLinden and S. A. Klein, Simulation of an absorption heat pump solar heating and cooling system, Solar Energy 31, 473-482 (1983). 18. E. Brousse, B. Clandel and J. R. Martine, Solar operated water ammonia absorption heat pump for air conditioning modelling and simulation, Appl. Energy 14, 131-142 (1983).