Performance of an HCFC22-based vapour absorption refrigeration system

[ ~ V T T E ~ R W Q R T H ~"~E I N E M A N N

Int. J. Refrig. Vol. 18, No. 7, pp. 465-476, 1995 Elsevier Science Ltd and IIR

Printed in Great Britain 0140-7007/95/$10.00 + .00

Performance of an HCFC22-based vapour absorption refrigeration system

M. Fatouh* and S. Srinivasa Murthy Refrigeration and Airconditioning Laboratory, Department of Mechanical

Engineering, Indian Institute of Technology, Madras 600 036, India Received 14 October 1994; revised 6 Apri l 1995

A single-stage vapour absorption refrigeration system (VARS) is tested with monochlorodifluoromethane (HCFC22) as refrigerant and different absorbents: dimethylether of tetraethylene glycol (DMETEG) and dimethyl acetamide (DMA). The influence of generator temperatures in the range 75-95 °C, which represents low-grade heat sources, is studied. Cooling water temperatures were varied between 20 and 30 °C. Two cases of cooling water flow paths are considered, i.e. water entering either absorber or condenser, which are connected in series. For HCFC22-DMETEG, COP values in the range 0.2-0.36 and evaporator temperatures between 0 and 10 °C are obtained. For HCFC22-DMA, COP values in the range 0.3-0.45 and evaporator temperatures between -10 and 10 °C are obtained. It is observed that HCFC22-DMETEG can work at lower heat source temperatures than HCFC22-DMA. However, at the same operating conditions HCFC22-DMA is better from the viewpoints of circulation ratio and COP. Experiments also show that at low heat source temperature, cooling water temperature has strong influence on circulation ratio but does not affect COP significantly. Preferably, cooling water should first flow through the condenser and then through the absorber in order to achieve improved overall performance. (Keywords: vapour absorption refrigeration system; HCFC22; DMA; DMETEG; experiments; cooling water flow path; COP; circulation ratio)

Performance d'un syst me frigorifique/t absorption de vapeur fonctionnant avec le HCFC22

On a test~ un systkme frigorifique mono~tag~ it absorption de vapeur fonctionnant avec le monochlorodi- fluoromOthane (HCFC22) comme frigorigkne et diff&ents absorbants comme le Other dimOthylique de glycol t~tra~thylOnique ( DMETEG ) et l'acdtamide dim~thylique (DMA ). On a ~tudid l'influence des temp&atures de g~n&ateurs allant de 75 it 95 °C (sources de chaleur de moindre importance). Les tempOratures de l'eau de refroidissement se sont dchelonnOes entre 20 et 30°C. On a testd deux possibilitds d'entr~e de l'eau de refroidissement: soit par l'absorbeur, soit par le condenseur; tout les deux dtant liOs en s&ie. Pour le couple HCFC22-DMETEG, on a obtenu des COP entre 0,2 et 0,36, ainsi que des tempdratures d'~vaporation entre 0 et 10 °C. Pour le couple HCFC22-DMA, on a obtenu des COP entre 0,3 et 0,45, ainsi que des temperatures d'Ovaporation entre - 1 0 et 10 °C. On a constat~ que le couple HCFC22-DMETEG peut fonctionner it des temp&atures plus basses que celles auxquelles fonctionne le couple HCFC22-DMA. Cependant, aux m~mes conditions de fonctionnement, le HCFC22-DMA est avantagd des points de rue du taux de circulation et du COP. Les expOriences ont dgalement montr~ que la temp&ature de l'eau de refroidissement poss~de, it une temperature basse de source de chaleur, une forte influence sur le taux de circulation, mais peu d'effet sur le COP. L'eau de refroidissement devrait, de preference, traverser le condenseur, puis l'absorbeur, afin de r~aliser une amelioration de la performance globale. (Mots clrs: Systrme frigorifique de production de froid; HCFC22; DMA; DMETEG; exprriences; chemin d'rcoulement de l'eau de refroidissement; COP; taux de circulation)

Vapour absorption refrigeration systems (VARS) have attained great importance in recent years, for two main reasons. Firstly, VARS can utilize low potential heat sources like waste heat, solar energy, low-pressure steam, etc., and thus can contribute significantly to energy conservation. Absorpt ion systems, which can use a variety of environmentally friendly fluid combinations,

*Permanent address: Mechanical Power Engineering Department, Faculty of Engineering & Technology, E1-Mattaria, Cairo, Egypt.

can contribute to overcoming the hazards caused by con- ventional CFCs. Moreover, thermal energy that would otherwise have been let out to warm up the atmosphere can be used to operate the absorption systems.

The drawbacks of the well-known ammon ia -w a t e r combinat ion are toxicity, high solution heat capacity, the need for rectification, and corrosivity to copper. The main problems associated with the other well-known solution of water- l i th ium bromide are the possibility of crystallization, the limitation of evaporator temperatures

465

466 M. Fatouh and S. Srinivasa Murthy

Nomenclature

COP CR h M P O T V W X

Coefficient of performance Circulation ratio Specific enthalpy (kJ kg- 1) Mass flowrate (kg s -1) Pressure (bar) Heat transfer rate (kW) Temperature (K) Specific volume (kg m -3) Pumping power (kW) Weight fraction of HCFC22 in solution (kgr/kgss)

Subscripts

1-24 Measurement locations in Figure 1

a Absorber act Actual c Condenser cw Cooling water e Evaporator eq Equilibrium g Generator h Solution heat exchanger hw Hot water p Pump r Refrigerant rw Refrigerated water ss Strong solution th Thermodynamic ws Weak solution

above the freezing point of water, corrosivity, high viscosity and vacuum operation. In view of the low vapour pressure of water, small pressure drops can only be permitted across the components of the system.

HCFC22 (ozone depletion potential=0.05) has now been included in the refrigerants to be phased out by the year 2030 at the Copenhagen Meeting ~ . However, it still occupies the position of a transitional fluid and is of great importance to developing countries. The use of HCFC22 in VARS can not only contribute to a reduction of global warming due to the consumption of waste heat but also can help in widening the application area of HCFC22 in the transition period.

With the use of the fluorocarbon-based refrigerant- absorbent combinations in VARS many of the drawbacks of ammonia-water and water-lithium bromide systems can be overcome. Numerous investigations 2-9 on halogenated hydrocarbons as refrigerants and organic solvents as absorbents have shown the mixture HCFC22-DMETEG to be one of the most promising combinations, as it yields high difference in boiling points between refrigerant and absorbent (needing no rectification), high solubility, safety, chemical stability and low generator cut-off temperature l°.

Many investigators 11-14 have analysed HCFC22- DMETEG absorption refrigeration systems based on thermodynamic and heat and mass transfer criteria to predict the performance characteristics. However, only one experimental work on HCFC22-DMETEG has been reported by Bulck et al. 15. At evaporator temperatures of 0 to -5 °C, condenser temperatures of 30-40 °C and at generator temperatures of 85-95 °C, their experiments yielded COP values in the range 0.25-0.30.

From rigorous comparative analyses on HCFC22- based VARS, the authors 1°'16-19 have found that among a wide variety of possible organic solvents, dimethyl acetamide (DMA) yields the highest COP. Another study highlighting HCFC22-DMA as a promising VARS fluid is presented by Borde and Jelinek ~°.

In the light of the above, experiments on a single-stage vapour absorption refrigeration system were carried out with HCFC22-DMETEG and HCFC22-DMA solutions as working fluids. The influence of cooling water temperature, heat source temperatures and evaporator temperatures are discussed. The influences of cooling

water flow direction through the absorber and condenser, which are connected in series, is also discussed. A comparison of the performances of HCFC22- DMETEG and HCFC22-DMA is made.

Description of the set-up

Figure 1 gives a schematic diagram of the single-stage VARS. The set-up consists of the VARS, heat source simulator, cooling load simulator, cooling water thermo- stat, instrumentation, and safety and operational controls. Table 1 gives the specifications of the main components of the VARS.

The VARS consists of two working fluid circuits. The primary circuit, in which pure refrigerant flows, consists of the condenser, expansion valve and evaporator. The secondary circuit, in which solution circulates, consists of the absorber, solution pump, solution heat exchanger, generator and expansion valve. The heat source simulator consists of a hot water tank, pump, flowmeter, heat exchanger, control valves and electric heaters to simulate the low-potential energy source. The cooling water source consists of a constant-temperature cooling bath, pump, flowmeter and control valves to supply cooling water to condenser and absorber in the desired direction. The condenser and absorber are connected in series with respect to the cooling water path and the direction of cooling water flow can be reversed. The temperature of the refrigerated water at the inlet to the evaporator may be maintained constant at a desired value using an electric heater, which simulates the refrigeration load.

The locations of various sensors are indicated in Figure 1. Precalibrated copper-constantan thermo- couples and inductive pressure gauges are used. Magnetic rotameters are used to monitor the flowrates of the liquid refrigerant and the weak solution, while turbine flowmeters are used to measure the flowrate of hot water and cooling water. The flowrate of refrigerated water is measured directly by observing the time to collect a known quantity.

Owing to the absence of an instrument for direct on- line measurement of concentration, an indirect technique is adopted. This was achieved by isolating a small quantity of solution in two vessels, which are located in parallel in both the strong and weak solution lines, as

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shown in Figure 1. Magnetic stirrers are used in both vessels to prevent stratification, and also sufficient time is allowed for the solution trapped in both vessels to reach equilibrium condition. The pressure and temperature in both vessels are measured and the concentration is calculated from the P - T X data l°.

Experimental procedure To start with, hot water at a predetermined temperature is circulated through the generator. The solution pump is then started to pump the strong solution to the generator. The weak solution level and pressure in the generator are observed. When the solution level in the shell reaches about two-thirds, at which condition all the pipes are submerged in the solution, the weak solution line valves are opened and the flowrate is regulated to maintain a constant level in generator.

When the pressure in the generator becomes greater than that in the condenser, the valve between the generator and condenser is opened to allow the refrigerant vapour to the condenser. Then cooling water in a series piping arrangement at the desired temperature is circulated through the absorber and condenser. The level of liquid refrigerant in the refrigerant receiver is noted.

Now the water to be refrigerated is allowed to pass through the evaporator and then liquid refrigerant is admitted through a capillary tube to the evaporator. The vapour line between the evaporator outlet and the absorber inlet is opened, admitting the refrigerant vapour into the absorber•

Throughout the experimental programme, the water flowrates in external circuits are kept constant at the design values, i.e. 0.35kgs -j for cooling water in the condenser-absorber circuit, 0.3 kgs -1 for hot water in the generator circuit, and 0.1 kgs I for refrigerated water in the evaporator circuit.

The system is run till steady-state conditions of pressure, temperature, flowrate and liquid levels are reached. The constancy of pressure and temperature readings, and also of liquid levels in the refrigerant receiver and generator, indicate that the system has attained steady-state conditions. Pressures, temperatures and flowrates at various locations are recorded.

As mentioned earlier, two different sets of experiments were done with two working fluids, namely HCFC22-DMETEG and HCFC22-DMA. The first set of experiments were with DMETEG as absorbent. After all the tests were concluded, generation and condensation of HCFC22 were continued to recover as much of the refrigerant as possible. Then all the remaining weakened solution was pumped into a recovery tank. The system was evacuated and flushed three times with DMA to remove any traces of DMETEG. The system was again evacuated and charged with fresh DMA. Charging of refrigerant is done at the absorber by circulating the solution through the absorber and removing the heat of absorption through the cooling water• Since the quantity of absorbent is known, the required initial concentration is achieved by weighing the refrigerant charge. The concentration is cross-checked by the method described earlier.

468 114. Fatouh and S. Srinivasa Murthy

Table 1 Specifications of main components of VARS

Tableau 1 Caract&istiques des composants principaux du systbmes frigorifique h absorption de vapeur

Condenser Type Shell and tube Fluid circuit HCFC22 in shell side and cooling

water in tube side Shell Cast iron, 15 cm (ID) Tube Copper, ID = 1.58 cm,

O D = 2 . 1 3 c m Length of tube 70 cm Number o f tubes 32 Number o f passes 4 Heat transfer area 1.5 m 2

Evaporator Type Tube in tube Fluid circuit HCFC22 in the annulus and

refrigerated water in the inner tubes

Material Mild steel Outer tube size ID = 3.51 cm, OD = 4.22cm Inner tube size OD = 2.67 cm, ID = 2.09 cm Number o f tubes 10 Length o f each tube 90 cm Heat transfer area 0 .6m 2

Solution pump Type Diaphragm pump Stroke length 40 mm Plunger diameter 40 mm Flowrate 2001 h - l (max) Discharge pressure 25 kg cm-" (max)

Generator Type Shell and tube Fluid circuit Solution in shell side and hot water

inside tubes Shell diameter 33.65 cm Tube material Mild steel Nominal size ID = 1.58 cm, OD = 2.13 cm Total number of tubes 72 Number of passes 4 Length of each tube 1.15 m Heat and mass transfer area 5.54m 2

Absorber Type Falling film outside vertical tubes Fluid circuit Solution in the shell side and

cooling water inside the tubes Shell Cast iron, 40 cm (ID) Tube material Mild steel Number o f tubes 30 Height of the tube ID = 2.09 cm, OD = 2.67 cm Length o f the tube 132 cm Number of passes 1 Heat and mass transfer area 3.32 m 2

Solution heat exchanger Type Tube in tube Fluid circuit Weak solution in the annulus and

strong solution in the inner tubes Material Mild steel Outer tube size ID -- 3.51 cm, OD = 4.22 cm Inner tube size ID = 2.09 cm, OD = 2.67 cm Number of tubes 24 Length o f each tube 167.6 cm Heat transfer area 3.38 m z

Performance calculations

Heat quantities

Heat quantities calculated from the measured temperature, pressure and mass flowrates of the working fluids are designated as thermodynamic values. Thus, thermodynamic heat quantities represent values internal to the

system as they are estimated from working fluid properties. Actual heat quantities are determined based on the measured flowrate and temperature rise (or drop) of the water passing through each component. Hence actual heat quantities represent the values external to the system. It should be mentioned here that all numerical subscripts in the following sections are with reference to the measurement locations indicated in Figure 1.

From the measured values of pressure, temperature and concentration of the working fluid, the inlet and outlet specific enthalpy values of each component have been calculated based on the thermodynamic properties of HCFC22, HCFC22-DMETEG and HCFC22-DMA pairs ~°a6. Then, mass and energy balances are made at each component to compute the thermodynamic heat quantities as follows:

Heat supplied at the evaporator:

Qe,th = M r ( h 6 - h5) (1)

Heat rejected at the condenser:

Qc,th - - M r ( h 2 - h3) (2)

Heat rejected at the absorber:

Qa,th = Mrh6 + M w s h l 6 - M s s h 7 (3)

where

Mss = Mws + Mr (4)

Generator heat input:

Qg,th ----- Mlhl + M13h13 - M12h12 (5)

Recovered heat in the solution heat exchanger is calculated from the strong solution stream, which extracts heat from the weak solution stream, as follows:

ah , t h -~ M s s ( h l l - h l0 )

The pumping power to calculated by

Wp,th --'= MssVs(P9 - / °8 )

pump

(6)

the strong solution is

(7)

where V s is the specific volume of strong solution at the inlet of the solution pump.

Actual heat transfer rates are determined from the measured water flowrate and temperature difference through each component as follows:

Qa,act = M~w(hj8 - h i7 )

Qg,act = Mhw(h23 - h24)

Qe,act = Mrw(h21 - h22)

Oc,act = Mcw(h20 - h i9 )

(8) (9)

(10) (11)

Performance characteristics

Two important performance characteristics of VARS are circulation ratio (CR) and coefficient of performance (COP). CR is defined as the ratio of the mass flow of the strong solution to that refrigerant, i.e.

CR = M~S (12) Mr

Since the difference in boiling points of refrigerant and absorbent is high, pure refrigerant is considered to leave the generator. A mass balance of the absorber or

An HCFC22-based vapour absorption refrigeration system 469

generator gives

CR - 1 - Xws ( 1 3 ) Xss-Xws

The thermodynamic circulation ratio is calculated by assuming equilibrium conditions at the outlets of absorber and generator:

1 -- )(13,eq (14) C R t h - X7,e q _ •13,eq

where Xv,eq and Xl3,eq are the equilibrium concentrations at absorber and generator exits respectively obtained from P - ~ X data ~° using the measured values of temperature and pressure.

The actual circulation ratio is estimated from measured values of refrigerant and weak solution flowrates as

C R a c t __ M r + Mws (15) M r

The COP is defined as the ratio of the cooling load at the evaporator (Qe) to the energy input at the generator (Q~). As the solution pump input is very small, it has been neglected in the calculation of the COP. Therefore

ae.th C O P t h - (16)

Qg,th The actual COP is calculated by means of the actual heat supplied at generator and evaporator:

C O P a c t _ Qe,act (17) Qg,act

R e s u l t s a n d d i s c u s s i o n

The following ranges of operating conditions could be studied depending on the limitations of the experimental set-up: 1. heat source temperatures between 75 °C and 95 °C to represent low-grade energy sources (higher temperatures were not attempted as the heat transfer fluid was water at atmospheric pressure); 2. cooling water temperatures in the range 20-30°C, which covers typical ambient conditions.

Cooling water through the condenser and absorber may flow in parallel or in series. Richter and Schumacher 21 have reported that usually a parallel layout of cooling water supply through condenser and absorber is adopted in industrial ammonia-water absorption plants. However, Kouremenos et al. 22 have discussed a compound ammonia-water/water-lithium bromide absorption refrigeration system in which cooling water flows in series from absorber to condenser. Herold and Radermacher 23 have mentioned that cooling water first goes to the condenser then to the absorber in a series piping arrangement for aqueous ternary hydroxide working fluid in their absorption heat pump. Recently, the authors 19 have made a detailed comparative analytical study of cooling water flow paths in VARS operating on HCFC22 with DMF, DMA and DMETEG as absorbents.

It is clear from the above literature that the direction of the cooling water path depends on the kind of

working fluid and the operating conditions, where both heat of absorption (latent heat of refrigerant) and heat of mixing play an important role. It is well known that parallel pipe connections for cooling water through the absorber and condenser result in better performance but demand large quantities of cooling water. However, in most practical systems it is customary to have a series connection for cooling the two components. Because of this, and also because of the limitation of the circulation pump in the experimental set-up, the case of parallel connection is not considered in the experimental study. Therefore in the present work a comparative study is made for two cases of cooling water pass in series as follows:

Case A: cooling water first goes to the absorber then to the condenser; and

Case B: cooling water flows first through the condenser then through the absorber.

With reference to Figure 1, various temperatures in this discussion are defined as follows:

1. heat source temperature T~: hot water temperature at the inlet of the generator, T23: 2. generator temperature Tg: temperature of weak solution at the generator outlet, T 13; 3. absorber temperature Ta: temperature of strong solution at the exit of the absorber, TT; 4. condenser temperature To: saturation temperature of refrigerant at the average pressure at the inlet and outlet of the condenser; 5. evaporator temperature Te: saturation temperature of refrigerant at the average pressure at the inlet and outlet of the evaporator; and 6. cooling water temperature Tcw: temperature of cooling water at the inlet of either the absorber or the condenser.

Results for HCFC22-DMETEG system

Effects of heat source temperature. When water flowrates are maintained constant and heat source temperature is changed, the operating temperatures of major components (generator, absorber, condenser and evaporator) adjust themselves to new values. It is clear from Figure 2 that operating temperatures at all components increase with heat source temperature. This is because of increases in the temperatures of both weak solution and refrigerant vapour at the respective outlets of generator. When the temperature of the refrigerant vapour leaving the generator and entering the condenser increases, with the cooling water flowrate through the absorber and condenser being kept constant, the condenser pressure and hence the condenser temperature increase. Because of the increase of the weak solution temperature at the outlet of the generator, the weak solution temperature at the inlet of the absorber increases and therefore the strong solution temperature at the absorber outlet also increases. This causes the evaporator pressure, i.e. the evaporator temperature, to increase.

The variations in concentration of strong and weak solution, circulation ratio and refrigerant mass flowrate are given in Figure 3. The weak solution concentration decreases at a faster rate than that of the strong solution


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Figure 2 Performance of HCFC22 DMETEG system: effects of heat source temperature on operating temperatures

Figure 2 Performance du systbme HCFC22-DMETEG: effets de la temperature de la source de chaleur sur les temperatures de fonctionnement

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Figure 4 Performance of HCFC22-DMETEG system: effects of heat source temperature on COP and heat quantities

Figure 4 Performance du systkme H C F C 2 2 - D M E T E G : effets de la tempOrature de la source de chaleur sur le C O P et les quantit#s de chaleur

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Figure 3 Performance of HCFC22-DMETEG system: effects of heat source temperature on concentration, circulation ratio and refrigeration mass flowrate

Figure 3 Performance du systbme HCFC22-DMETEG: effets de la tempOrature de la source de chaleur sur la concentration, le taux de circulation et de d~bit massique du frigorigkne

concentration when the heat source temperature increases. Hence, when the temperature and flowrate of the cooling water are held constant, the concentration spread (Xss -Xws) increases with heat source temperature, causing the refrigerant mass flowrate to increase and the circulation ratio to fall. As a result, both the strong and weak solution mass flowrates decrease. It is also seen that the actual circulation ratio is higher than that from thermodynamic analysis wherein equilibrium conditions at the exits of the generator and absorber are assumed. In practice, the actual strong solution concentration is usually lower than the equilibrium value, and the actual concentration of the weak solution is higher than equilibrium conditions. Therefore the actual concentration spread is lower than that estimated from thermodynamic analysis.

The variations of the coefficient of performance and

heat quantities at the main components are plotted in Figure 4. All heat quantities increase with heat source temperature. However, the generator heat input increases more sharply than that at other components. It is observed that the specific enthalpy of the superheated refrigerant vapour at the outlet of the generator (hi), that of the saturated vapour at the outlet of the evaporator (h6) and also that of the saturated liquid at the exit of the condenser (h3) are increased. Therefore, both cooling effect (h 6 - h s ) and enthalpy difference across the condenser (h 2 - h3) increase with heat source temperature. It is already known from Figure 3 that the refrigerant mass flowrate increases while both the strong and weak solution mass flowrates decrease with heat source temperature. Because of this, the thermodynamic cooling capacity (Qe,th) as given by Equation (1) increases. Also, owing to the combined effect of increases in enthalpy difference (h2 - h3) across the condenser and refrigerant mass flowrate, the condenser heat load increases as governed by Equation (2). It may be noted that the specific enthalpies of both the weak and strong solutions at the inlet and outlet of absorber respectively, and that of the strong solution at the inlet of the generator, are increased only slightly, while that of the weak solution at the generator outlet is increased significantly. Since the enthalpies of the weak and strong solutions at the inlet and outlet of the absorber are increased slightly while the mass flowrates of the weak and strong solutions decrease, the value of (mwshl6- mssh7) in Equation (3) increases by a small value. Hence the thermodynamic heat rejected at the absorber increases. In addition, the increases in cooling effect (latent heat) and refrigerant mass flowrate also cause the heat of absorption to increase. It is clear from Equation (5) that the main reason for the increase in thermodynamic heat input at the generator is the increase of heat associated with the refrigerant stream because of the increase in both the mass flowrate and the enthalpy of the superheated vapour of the refrigerant, while there is a small influence from the strong and weak solution streams. The decreases in both solution heat exchanger duty and pumping power are mainly due to


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Figure 5 Performance of H C F C 2 2 - D M E T E G system: effects of heat source temperature on operating temperatures

Figure 5 Performance du systbme HCFC22-DMETEG: effets de la tempOrature de la source de ehaleur sur les tempdratures de fonctionnement

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Figure 6 Performance of H C F C 2 2 - D M E T E G system: effects of heat source temperature on evaporator temperature

Figure 6 Performance du systbme HCFC22-DMETEG: effets de la temperature de la source de chaleur sur la temperature d'Ovaporation

the decrease of circulation ratio with increase in heat source temperature.

It is seen that the thermodynamic pumping power varies between 5 and 10% of the thermodynamic heat input at the generator. Hence it is neglected in the COP calculations. Moreover, the pump used here was what was available, and hence was not an optimal one.

As mentioned earlier, the actual heat quantities are calculated from the measured water flowrate and temperature change through each component. It is seen that the actual heat input at the generator is more than that estimated from thermodynamic analysis owing to the thermal losses to ambient through insulation. The actual quantities of heat rejected at both the absorber and condenser are less than the corresponding thermodynamic values because of heat losses to the surroundings. It is observed that the difference between the thermodynamic and actual heat quantities is higher at the absorber than that at the condenser. This is due to the large external surface area of the absorber for heat losses to take place. It is also seen that the COP decreases as the heat source temperature increases. This is due to the large increase in heat input at the generator while the cooling load at the evaporator is slightly increased. The actual COPac t values are obviously lower than the corresponding COPth values.

Effects of cooling water temperature and flow path. Ex- periments are conducted at different cooling water temperatures for the two cases of cooling water paths at different heat source temperatures keeping constant flowrates in the hot water, cooling water and refrigerated water circuits. The results are presented in Figures 5-8 as function of heat source temperatures.

Figure 5 reveals that for Case B, i.e. when cooling water flows first through the condenser then the absorber, there is a decrease in condenser temperature because the cooling water temperature at the inlet to the condenser is lower than that for Case A. Also, since the cooling water temperature at the inlet to the absorber is higher, the absorber temperature is higher. The increase

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Performance of H C F C 2 2 - D M E T E G system: effects of heat source temperature on concentration, circulation ratio and refrigeration mass flowrate

Figure 7 Performance du systkme HCFC22-DMETEG: effets de la tempdrature de la source de chaleur sur la concentration, le taux de circulation et le ddbit massique du frigorigbne

in absorber temperature causes the evaporator temperature to increase.

The variations of evaporator temperature with heat source temperature at two temperatures and directions of cooling water are shown in Figure 6. For a given heat source temperature as well as cooling water temperature, it is seen that the evaporator temperature is slightly higher for Case B than for Case A. This suggests that to achieve low temperatures at evaporator temperatures, the cooling water path should follow Case A.

Figure 7 makes a comparison of the variations in concentration of both strong and weak solutions, circulation ratio and refrigerant mass flowrate for two directions of cooling water path. It can be seen that a high concentration spread can be obtained for Case B, resulting in a lower circulation ratio and higher refrigerant mass flowrate. Particularly, from the viewpoint


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- Case Tc.=25.0 +_

0 4

( ~ ~ ~

75 ' 8 '0 ~ w ~ r~ 8 '5 9~ 0 r T ~ r ~ 9 '5 ' 1 ' 0 0

Heat sou rce t e m p e r a t u r e ( % )

Figure 8 Performance of H C F C 2 2 - D M E T E G system: effects of heat source temperature on COPae t

Figure 8 Performance du systbme H C F C 2 2 - D M E T E G , effets de la tempdrature de la source de chaleur sur le COP~c ~

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05

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- C}as ,P,,

[c~( °~ )

m u n r o • 2 b

75 80 85 90 95 !(>0 Heat sou rce t e m p e r a t u r e ( ° ( )

Figure 10 Performance of H C F C 2 2 - D M E T E G system: effects of heat source temperature on COP~ t

Figure 10 Performance du systkme HCFC22-DMETEG: effets de la temperature de la source de chaleur sur le COP~ct

2 0 HCFC22-DMETEO

- Case

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O" 1 T , , , , , I , w T m m ~ v T ~ F ~ - - m , , , , , I ' I ' , ' r ' ' ' 75 80 85 90 95 100

Heat sou rce t e m p e r a t u r e ( C )

Figure 9 Performance of H C F C 2 2 - D M E T E G system: effects of heat source temperature on circulation ratio

Figure 9 Performance du systbme H C F C 2 2 - D M E T E G : effets de la temperature de la source de chaleur sur le taux de circulation

of low CR, Case B shows significant advantage over Case A.

A comparison of COPac t for the two directions of cooling water path is shown in Figure 8. It is clear that a high COPact can be achieved with Case B. This is mainly due to the slight decrease in Qg and also the slight increase in Q~, both of which contribute to an increased COP for Case B.

Figure 9 shows the variation of circulation ratio for the two cooling water passes at different temperatures. The circulation ratio increases with the cooling water temperature. As the cooling water temperature increases, the condenser pressure and hence the generator pressure increase. This causes the weak solution concentration to increase for a given heat source temperature. However, the change in strong solution concentration is negligible owing to changes of both the evaporator and absorber temperatures. Thus the concentration spread decreases

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0 2 4 6 8 ,

F v u p o r a t o r t e m p e r a t u r e I, ~ , )

Figure 11 Performance of HCFC22-DMETEG system: effects of evaporator temperature on concentration, circulation ratio and refrigerant mass flowrate

Figure 11 Performance du systbme HCFC22-DMETEG: effets de la temperature de l'~vaporateur sur la concentration, le taux de circulation et le d~bit massique du frigorigbne

as the cooling water temperature increases, causing the circulation ratio to increase. It may be noted that the cooling water temperature exerts a stronger influence on the circulation ratio for Case A than for Case B, particularly at low heat source temperatures.

From a comparison of COP at different temperatures and passes for cooling water shown in Figure 10, it is clear that high COP values can be obtained with either low cooling water temperature, or the Case B path, in which cooling water enters first the condenser and then the absorber. However, at low heat source temperatures the cooling water temperature does not exert a significant influence on COP. It is also clear that COPs in the range 0.2-0.36 can be obtained over the entire range of heat source temperature from 75 to 95 °C and cooling water temperatures between 20 and 30 °C.

An HCFC22-based vapour absorption refrigeration system

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J 1 Q h " [ ..... 2b ' : ) +~ 0 5 ° C P 0 4 5 I _ ~COP ~

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Figure 12 Performance of HCFC22-DMETEG system: effects of evaporator temperature on COP and heat quantities

F i g u r e 12 Performance du systkme HCFC22-DMETEG: effets de la tempOrature de 1Uvaporateur sur le COP et les quantit~s de chaleur

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Figure 13 Performance of HCFC22 DMA system: effects of heat source temperature on operating temperatures Figure 13 Performance du systbme HCFC22-DMA: effets de la temperature de la source de chaleur sur les temperatures de fonctionnement

Effects of evaporator temperature. For these studies, Case A of cooling water path is alone chosen because, as discussed with reference to Figures 5 and 6, low evaporator temperatures could be achieved by this config- uration.

The expected variations of both strong and weak solution concentrations, circulation ratio and refrigerant mass flowrate are revealed in Figure 11. Both the strong and weak solution concentrations increase with evaporator temperature. The increase in strong solution concentration is due to the increase of evaporator temperature, while that in weak solution concentration is because of the increase in condenser pressure at constant generator temperature. It is also clear that as the evaporator temperature increases the concentration spread increases, thereby causing the circulation ratio to decrease and the refrigerant mass flowrate to increase.

Figure 12 gives the variations of COP and heat

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85 90 95 / o - H e a ! s o u r e f e r n p e r s f j r e , k , )

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Figure 14 Performance of HCFC22-DMA system: effects of heat source temperature on concentration, circulation ratio and refrigerant mass flowrate Figure 1 4 Performance du systbme HCFC22-DMA: effets de la temprrature de la source de chaleur sur concentration, le taux de circulation et le d~bit massique du frigorigbne

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Figure 15 Performance of HCFC22-DMA system: effects of heat source temperature o n C O P a c t and heat quantities Figure 15 Performance du systbme HCFC22-DMA: effets de la temprrature de la source de chaleur sur le COPac t et les quantites de chaleur

quantities of the main components with evaporator temperature. The generator heat load is nearly constant, and the heat loads at the evaporator, condenser and absorber increase while the solution heat exchanger duty decreases. The decrease in solution heat exchanger duty is mainly due to the decrease in circulation ratio with increase in evaporator temperature, as already seen in Figure 11. As the evaporator temperature increases, there is a small increase in both the refrigerant mass flowrate and the cooling effect, while the change of enthalpy difference at the inlet and outlet of the condenser is not significant. Therefore, the thermodynamic cooling capacity as well as the heat rejected at the condenser condition increase. Owing to the increase of both refrigerant mass flowrate and cooling effect, the heat of

4 7 4 M. Fatouh and S. Srinivasa Murthy

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Figure 16 Performance of HCFC22-DMA system: effects of heat source temperature on operating temperatures Figure 16 Performance du systbme HCFC22-DMA." effets de la temperature de la source de chaleur sur les temperatures de fonctionnement

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H e a t s o u r c e t e m p e r a t u r e (°C)

Figure 18 Performance of H C F C 2 2 - D M A system: effects of heat source temperature on COPac t

Figure 18 Performance du systOme HCFC22-DMA." effets de la temperature de la source de chaleur sur le COPac t

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Figure 17 Performance of HCFC22-DMA system: effects of heat source temperature on concentration, circulation ratio and refrigerant mass flowrate Figure 17 Performance du systbme HCFC22-DM,4: effets de la tempOrature de la source de chaleur sur la concentration, le taux de circulation et le d~bit massique du frigorigbne

2O ¢~ HCFC22-DMA cO

Case Case

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o

.~ d m_m@im 25 0 t A_ A_A A_A 20

70 75 80 88 90 95 100

H e a t s o u r c e t e m p e r a t u r e (°C)

Figure 19 Performance of H C F C 2 2 - D M A system: effects of heat source temperature on circulation ratio and refrigerant mass flowrate

Figure 19 Performance du systkme H C F C 2 2 - D M A . effets de la temperature de la source de chaleur sur le taux de circulation et le ddbit massique du frigorigbne

absorption increases. This causes the thermodynamic heat rejected at absorber condition to increase. It can be seen that the COP increases with evaporator temperature. The reason for this is that the heat input at the generator is nearly constant while the cooling capacity at the evaporator increases. The actual COP is less than the thermodynamic COP owing to the increase of actual heat input at the generator.

Results for H C F C 2 2 - D M A system and comparison with HCFC22-DMETEG system

Figures 13-24 show the results obtained when the VARS is operated with H C F C 2 2 - D M A solution. Here, for the sake of clarity, the thermodynamic values of the parameters are omitted and only the actual values are

presented. The water flowrates through the external circuits and solution flowrates were kept the same as those for the H C F C 2 2 - D M E T E G system. It is observed that the general trends of variation of all parameters are similar for both systems. From the results, it is possible to make a qualitative performance comparison between the systems.

The H C F C 2 2 - D M E T E G system can operate at lower heat source temperatures than the H C F C 2 2 - D M A system. Also, the H C F C 2 2 - D M E T E G system can perform with higher heat rejection temperatures at the condenser and absorber. At similar operating conditions, the H C F C 2 2 - D M A system operates at higher concentration levels and can also yield lower evaporator temperatures. As seen in Figure 25, the H C F C 2 2 - D M E T E G system requires higher circulation ratios and also its circulation ratio increases more steeply with decreasing


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Figure 211 Performance of H C F C 2 2 - D M A system: effects of heat source temperature on COP~c t

Figure 20 Performance du systbme HCFC22-DMA: effets de la temperature de la source de chaleur sur le COPac ~

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Figure 22 Performance of HCFC22-DMA system: effects of evaporator temperature on COPact and heat quantities

Figure 22 Performance du systbme HCFC22-DMA: effets de la temperature de l'Ovaporateur sur le COPact et les quantit~s de chaleur

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-,( 5 o 5 E ,<apo ra t c ' t e m p e r a t u r e (°C)

Figure 21 Performance of HCFC22-DMA system: effects of evaporator temperature on concentration, circulation ratio and refrigerant mass flowrate

Figure 21 Performance du systbme HCFC22-DMA: effets de la temperature de l'~vaporateur sur la concentration, le taux de circulation et le ddbit massique du frigorigkne

heat source temperature. Higher COPs of the HCFC22- DMA system are observed in Figure 26, owing to better refrigerant release at the generator at any given heat source temperature.

Conclusions

An experimental programme to study the performance of a VARS is described, leading to the following observations:

1. Vapour absorption refrigeration systems working on HCFC22-DMETEG or HCFC22-DMA solution can operate using low-temperature heat sources, in the range of about 75-95 °C.

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Figure 23 Performance of HCFC22-DMA system: effects of evaporator temperature on circulation ratio

Figure 23 Performance du systbme HCFC22-DMA: effets de la tempOrature de l'~vaporateur sur le taux de circulation

2. Improvements in COP and low circulation ratio are obtained when the cooling water flows first through the condenser then goes to the absorber (Case B). 3. Cooling water temperature does not significantly affect COP but it influences circulation ratio, particularly at low heat source temperatures. 4. At low heat source temperatures, the solution heat exchanger duty is the maximum heat load of the VARS's components. 5. The HCFC22-DMA system operates at lower CR and yields better COP than the HCFC22-DMETEG system at similar operating conditions. 6. Actual COPs in the range of 0.2-0.36 and evaporator temperature from about zero to 10 °C for HCFC22- DMETEG, and COPs in the range of 0.3-0.45 and evaporator temperature from about -10 to 10°C for HCFC22-DMA are obtained over the range of cooling water temperatures of about 20-30 °C and heat source temperature between 75 and 95 °C.


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T¢~= 20°C I - - T~= 25°C

r , , , , , i , ' , , m T ~ r : r q 5 0 5

Evapo ra to r t e m p e r a t u r e (°C)

Figure 24 Performance of HCFC22-DMA system: effects of evaporator temperature on COPac t

Figure 24 Performance du systbme HCFC22-DMA: effets de la temperature de l'~vaporateur sur le COPac t

20

15

0

0

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70 75 80 85 90 95 100 Heat s o u r c e t e m p e r a t u r e (°C)

Figure 25 Performance comparison between HCFC22-DMETEG and HCFC22-DMA systems

Figure 25 Comparaison des performances entre les systkmes HCFC22- DMETEG et HCFC22-DMA

References

1 Kuijpers, L. J. M. Copenhagen 1992: a revision or a landmark? Int J Refrig (1993) 16 210-220

2 Eiseman, B. J. Why refrigerant HCFC22 should be favored for absorption refrigeration A S H R A E J (1959) 1 45-50

3 Albright, L. F., Doody, T. C., Budez, P. C., Pluche, C. R. Solu- bility of refrigerant 1 l, 21 and 22 in organic solvents containing an oxygen atom ASHRAE Trans (1960) 66 423-433

4 Thieme, A., AIbright, L. F. Solubility of refrigerants 11, 21 and 22 in organic solvents containing a nitrogen atom and in mix- tures of liquids ASHRAE J (1961) 3 71-75

5 Kriebel, M., Loftier, H. J. Thermodynamic properties of the bin- ary system HCFC22/E181 Kalte-technik (1965) 17 266-271

6 Ehmke, H. J., Steinfle, F. Investigation of organic solvents for the refrigerant HCFC22 Clima 2000 (1985) 6 569-574

7 Borde, I. Development of absorption refrigeration and heat pump units with organic working fluids Absorption Heat Pumps Congress Paris (1985) 468-473

8 Matsuo, K., Kunugi, Y., Usui, S. Absorption heat pump with organic fluid pair Absorption Experts Meeting (1985) 385-398

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:2'.4 i

05 t

EL @ <-)02 i

0 J

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A

HC~C'22-DMA -- - - '~ - - HCFC22-DMETEO r~=2~( ; ± O.S°C

,AA,A Cose (a) J =_m==_w Cose (B)

O. 0 ~ ~ q ~ r r ~ ~ ~m~r r~T

7~, 75 80 85 90 95 100

t tea t s o u r c e t e m p e r a t u r e ( ~ )

Figure 26 Performance comparison bewteen HCFC22-DMETEG and HCFC22-DMA systems

Figure 26 Comparaison des performances entre les systbmes HCFC22- DMETEG et HCFC22-DMA

9 Kawamoto, K., Takata, Y., Shibuya, K. Absorption heat pump with HCFC22 and organic solvent Absorption Heat Pumps Congress Paris (1985) 399-414

10 Fatouh, M., Srinivasa Murthy, S. Comparison of HCFC22- absorbent pairs for absorption cooling based on P - T X data J Renewable Energy (1993) 3 31-37

11 Borde, I., Jellnek, M., Yaron, I. Continuous solar heated absorption cooling unit for industrial applications Int Sol Forum Hamburg (1979) Part II 511 521

12 Renz, M., Steimle, F. Comparison of thermodynamic properties of working fluids for absorption systems Proc IIR Commissions El, E2 Jerusalem (1982) 1-7

13 Borde, I., delinek, M. Development of absorption refrigeration units for cold storage of agricultural products lnt J Refrig (1987) 10 53-56

14 Kumar, S., Prevost, M., Bugarel, R. Comparison of various working pairs for absorption refrigeration systems: application of R21 and HCFC22 as refrigerants Int J Refrig (1991) 14 304-310

15 Buick, E. V. D., Trommelmans, J., Berghmans, J. HCFC22-E 181 solar absorption cooling installation Proc IIR Commissions E1 E2 Jerusalem (1982) 83-87

16 Fatouh, M., Srinivasa Murthy, S. Comparison of HCFC22- absorbent pairs for absorption cooling based on H - T - X data J Renewable Energy (1993) 3 262-271

17 Fatouh, M., Srinivasa Murthy, S. HCFC22 based vapour absorption refrigeration systems, Part lh Influence of component effectiveness Int J Energy Research accepted

18 Fatouh, M., Srinivasa Murthy, S. HCFC22 based vapour absorption refrigeration systems, Part II: Influence of component effectiveness Int J Energy Research accepted

19 Fatouh, M., Srinivasa Murthy, S. HCFC22 based vapour absorption refrigeration systems, Part IIh Effects of different absorber and condenser temperatures Int J Energy Research accepted

20 Borde, L Jelinek, M. Absorption heat pumps with organic refrigerant-absorbent fluid pairs Second Int Workshop on Research Activities on Advanced Heat Pumps Graz, Austria (1988) 89-101

21 Richter, K. H., Sehumacher, G. Design, operation and comparative evaluation of industrial ammonia-water absorption plant Linde Reports on Science and Technology (1974) 20 26-55

22 Kouremenos, D. A., Rogdakis, E., Antonopoulos, K. A. A high- efficiency, compound NH3/H20-H20-LiBr absorption refrigeration system Energy (1989) 14 893-905

23 Herold, K. E., Radermacher, R. Development of an absorption heat pump water heater using aqueous ternary hydroxide working fluid Int J Refrig (1991) 14 156-167

Documents

Performance of an HCFC22-based vapour absorption refrigeration system