gas-driven absorption heat-pump systems

~

Applied Thermal Engineering Vol. 16, No. 4, pp. 347-356, 1996 Elsevier Science Ltd 1359-4311(95)00085-2 Printed in Great Britain 1359-4311/96 $15.00 + .00

Pergamon

SOLAR/GAS-DRIVEN ABSORPTION HEAT-PUMP SYSTEMS M. Nguyen, S. B. Riffat

and D. Whitman*

Building Technology Institute, School of Architecture, University of Nottingham, University Park, Nottingham NG7 2RD, UK; and *British Gas, Gas Research Centre, Ashby Road, Loughborough, Leicestershire LE11 3QU, UK (Received 28 February 1995) Abstract--This paper describes the potential for the development of hybrid gas/solar-driven heat-pumps with the specific examples of a combined refrigeration and power-generation system and a sea-water purification system. Thermodynamic analysis has been performed on the systems using the waterammonia and the lithium-bromide-water combinations. A desalination plant of this type is shown to have scope for a higher water yield than conventional solar stills, and the refrigeration system shows high efficiency coupled with flexibility of operation. Keywords--Absorption cycle; heat pumps; hybrid system; solar energy; thermodynamic analysis. NOMENCLATURE COP R coefficient of performance (cooling) specific enthalpy (kJ kg- i ) h mass flow rate (kg s -l) m pressure in absorber (kPa) PA pressure in generator (kPa) e~ pressure in condenser (kPa) Pc pressure in evaporator (kPa) PE Qo generator heat input (kW) rA temperature in absorber (°C) temperature in generator (°C) ro rc temperature in condenser (°C) /'E temperature in evaporator (°C) T~ ambient temperature (°C) Greek letters turbine thermal efficiency (%) qT

1. I N T R O D U C T I O N

The concept of solar-powered refrigeration cycles is not new and several machines operating on this principle are now commercially available. However, there has been little research into the integration of solar power with natural gas. Such a system could remain operational when solar insolation is low, permitting its all-year-round operation and use in areas where solar energy alone is impractical. Many solar-powered domestic water-heating systems have provision for an electrical immersion heater as a back-up [1]. Use of a gas back-up would be more economical and utilising gas and solar power simultaneously would reduce the cost and size of solar collectors employed in solar-driven absorption systems. Vapour-compression systems have long been favoured in preference to absorption-type systems for refrigeration applications, due to their significantly higher performance. However, use of a gas/solar-driven system would reduce energy costs, making absorption systems more attractive on economic grounds. The use of absorption technology over vapour-compression technology is highly desirable from an environmental viewpoint. Generally, most vapour-compression systems use CFCs or HFCs as working fluids. These compounds are known to be highly damaging if leaked to the atmosphere and their manufacture and use will cease in the foreseeable future. Many of the alternative working fluids used in vapour-compression cycles lead to a reduction of performance and require substantial 347

348

M. Nguyen et

al.

modifications to the system [2, 3]. This, allied with the greater cost of alternative refrigerants, will no doubt improve the competitiveness of absorption machines. Currently, most vapour-compression machines are electrically powered. Electrical power generation and transmission efficiencies are low, typically of the order of 30%. Most power generation techniques are highly damaging to the environment, leading to acid rain, river pollution and the depletion of valuable fossil fuel resources. A gas/solar-driven absorption system would be less damaging as there are much lower emissions associated with natural gas than other fossil fuels. Indeed, natural gas could even be supplied from the decay of waste materials in landfill sites (biogas) where suitable. In this paper two applications of solar/gas-driven absorption heat pumps are considered, namely a combined refrigeration and power-generation system and a salt-water desalination system. Many other applications are possible, such as product concentration, drying and heating. A combined refrigeration and power system offers several advantages over a refrigeration system alone. Excess cooling capacity may be reduced and power generation increased, thereby allowing the system to run continuously at maximum efficiency. Electricity produced may be used to power other building services, such as lighting or ventilation, or it may be used to power-up an energy store to supplement the natural gas supply in periods of low solar insolation. A combined refrigeration and power-generation system would allow operation independently of an external power supply. In many areas of the world, power supply is intermittent or non-existent. Most absorption systems require power for control gear and a combined system would allow unencumbered operation. A suitable energy-storage strategy may be to pass excess electrical power to a catalytic electrolysis plant, producing hydrogen and oxygen. It is thought by many environmentalists and scientists alike that the use of hydrogen will replace some fossil fuels in the not too distant future, especially in developing countries [4]. This hydrogen could be catalytically burnt in the generator to supplement natural gas supplies when solar insolation is low or could be passed, along with oxygen, directly to a highly efficient fuel cell to produce power, or could be used as an alternative to fossil fuels in other applications. The benefits of an absorption sea-water desalination system are numerous and include the following: water extraction occurs at low temperatures, reducing problems of scaling and corrosion; no water pretreatment is required; the system is quiet and requires little maintenance. An absorption sea-water desalination system utilising the indirect freezing method [5] could allow much higher efficiencies than those associated with conventional multi-stage flash systems. 2. D E S C R I P T I O N Absorption systems utilise an absorber, a generator, a condenser, an evaporator and a solution pump. The energy required by the absorption system comes mainly from the heat supplied to the generator. However, a small amount of work is required by the solution pump. This is in contrast to conventional vapour-compression systems, which require shaft work for the compression process. Figure 1 illustrates the cooling cycle of the combined refrigeration and power-generation system. In the evaporator, the low-pressure refrigerant liquid evaporates, the process being endothermic. Heat is supplied to the evaporator by air or water; this chilled air or water is subsequently used for refrigeration. Refrigerant vapour then enters the absorber at low temperature and low pressure through line 2. The refrigerant vapour is then dissolved by the weak aqua-ammonia solution from line 8 in the absorber. Heat liberated during this process is removed using a water or air sink. The solution pump receives strong aqua-ammonia solution through line 3 and delivers it at high pressure to the generator via the heat exchanger through line 4. Note that the solution is a saturated liquid and hence the pump work is very small. The high-pressure solution entering the generator through line 5 is heated and some of the ammonia in solution is vaporised. The heat for this process could be provided by solar energy and a gas burner. The remaining weak aqua-ammonia solution exits through line 6 via the heat exchanger and to the absorber through a pressure-reducing valve between lines 7 and 8. A heat exchanger is included here as the solution in line 4 requires heating and the hot solution in line 6 would increase the absorber temperature if it were not cooled before

Solar/gas-driven absorption heat-pump systems I~

Solar Collector (Generator)

349

9,1'

J I / Condenser

5 Heat Exchanger

i v

E~

7 4

~N

;Qoo, Solution Pump

Absorber

Evaporator

LiBr/H20

Fig. I. Hybrid gas/solar energy heat-pump for refrigeration.

entry. High-pressure, high-temperature ammonia vapour leaving the generator is passed to the condenser through line 9. In the condenser the refrigerant vapour is cooled and condenses to a liquid. This liquid exits through a pressure-reducing valve in line 1 and into the evaporator. Figure 2 illustrates the combined refrigeration and power-generation cycle. The power-generation cycle is essentially a Rankine cycle with the Rankine condenser replaced by the absorber. High-temperature, high-pressure ammonia vapour leaving the generator may be passed to the turbine through line l0 instead of to the condenser. This vapour is expanded through the turbine to the lower pressure in line 2. The low-pressure vapour is then passed to the absorber. The turbine is coupled to an electrical generator, the electrical load may be other building services or a catalytic electrolysis plant in addition to the solution pump and control gear. Significant increases in the C O P R and qT may be achieved if the turbine and condenser operate at different temperatures and pressures. The power-generation system and refrigeration system ®

P I

9

Load Eletrical

~

Solar Collector (Generator)

~-~

I

t

Condenser Controller

"11

5 Heat Exchanger

Solution

Pump

Evaporator

Absorber LIBr/H20

Fig. 2. Hybrid gas/solar energy heat-pump for refrigeration and power generation.

350

M. Nguyen et

al.

share the same absorber, solution pump and heat exchanger but separate generators are provided for the two systems. An additional pressure-reducing valve is added to reduce the pressure of the solution entering the generator coupled to the condenser. The capital cost of this system would be greater than that illustrated in Fig. 2 but greater efficiencies would lead to lower running costs. The water-ammonia working pair has been chosen as ammonia has a high vapour pressure at modest temperatures. This is necessary to provide the large pressure difference across the turbine. It is also desirable to operate such systems at higher than atmospheric pressure to prevent leakage of air into the system, which would reduce the system's performance. Whilst this system has been analysed for the water-ammonia working pair, other suitable combinations may exist. These include the carbon-monoxide-acetone combination and the ammonia-sodium-thiocyanate combination. Figure 3 illustrates the salt-water desalination system. This system is also suitable for drying and concentrating processes. A weak brine solution (feed) enters the evaporator through a pressure-reducing valve located in line 1. Heat may be added to the evaporator from the absorber and/or the condenser. The low pressure in the evaporator causes the brine solution to boil and the pure water vapour produced exits through line 2. The strong brine solution exits through line 3 to the discharge pump. The pump returns the brine solution to atmospheric pressure before it is discarded. Pumping work is negligible as the solution is a saturated liquid and hence may be regarded as incompressible. The water vapour enters the absorber through line 2, where it is absorbed by the strong lithium bromide solution in line 9. The resulting weak lithium bromide solution passes to the solution pump via line 4, where its pressure is raised. Again the pumping work is very small. The weak solution passes from the pump to the heat exchanger through line 5, where it is preheated prior to entering the generator. Again this leads to better performance and lower absorber cooling load. The weak solution then enters the generator through line 6, where it is heated by a combination of solar energy and gas. The high temperature drives off some of the water from the lithium bromide solution as a vapour. This pure water vapour then passes through line 10 to the condenser, where it is cooled and condenses. The strong lithium bromide solution in the generator exits through line 7, passes through the heat exchanger and pressure-reducing valve, back to the absorber through line 9. A water-based working pair must be used in this system and the lithium-bromide-water combination has been chosen here. Other working pairs might also be suitable and these include water-caustic-soda-caustic-potash and water-potassium-formate. Ideally the absorbent should have a low vapour pressure to ensure that the water extracted contains few impurities, especially if the system is to supply potable fresh water. [~

S~IGermCe°llect°atror)/ I

I0!

/_ Condenser~

~

")

~

Burner

i:i:!.! . . . .

,

Solution Absorber Pump LIBr/H20

o.t

--~Conden~te

i ~ ~rge~

~

'

~

~ Brine

.

Evaporator

Fig. 3. Hybrid gas/solar energy heat-pump for desalination processes.

Solar/gas-driven absorption heat-pump systems

351

Higher performance may be possible using an absorption freezing-type method of desalination [5]. In this system evaporation of sea water occurs until the solution cools sufficiently to allow ice crystals to form. These pure ice crystals are then separated from the remaining brine solution. The ice is used to cool the condenser of the system, thus melting the ice and condensing water vapour from the generator. Potentially all these systems could be used for heating, utilising the absorber as a low-grade heat-source and the condenser as a high-grade heat-source. 3. R E S U L T S A N D D I S C U S S I O N The single-effect absorption systems for refrigeration and power generation were analysed thermodynamically using the water-ammonia working pair for a range of operating conditions. It was assumed that the heat-exchanger efficiency was constant at 80%, regardless of generator or absorber temperatures and flow rates. This simplification allows the temperature of the solution entering the generator to be calculated. The gas combustion efficiency is assumed to be 95% and the heat transfer from burner to generator is assumed to be 70%. For the power generation cycle the turbine isentropic efficiency is also assumed to be 70%. It is also assumed that the vapour at the turbine exit is dry and in some cases the turbine thermal efficiency is limited by this condition. Details of the methods for calculating the performance of absorption-type systems are given in many proprietary text books, including that by Stoecker and Jones [6]. Figure 4(a) and (b) shows the effect on the COPR and ~/T, respectively, of increasing the condenser temperature from 20 to 60°C for different generator temperatures. As condenser temperature increases, COPR decreases but ~/v increases. This occurs because the turbine and condenser are constrained to operate at the same pressure. Operating the turbine and condenser at different pressures allows different generator temperatures to be used, leading to further increases in the cooling performance and the power generation efficiency. Figure 5(a) and (b) shows the effect on the COPR and ~/T, respectively, of increasing the evaporator temperature from - 10 to 10°C for different generator temperatures. As the evaporator temperature increases, COPR increases but ~/T decreases. The optimum temperature will depend on the cooling and power loads at any given instant. Figure 6(a) and (b) shows the effect on COPR and ~/Tof increasing the absorber temperature from 10 to 50°C for different generator temperatures. As absorber temperature increases, both COPR and ~/Tdecrease. Ideally the absorber should operate at as low a temperature as possible. However, if the absorber is to be air-cooled, this will necessitate its temperature being several degrees above ambient. In mild climates this will not present a difficulty, but in hotter climates absorber temperatures of 40°C or more may be required. If a cool water source is available (e.g. ground water) then this problem may be alleviated. Table 1 shows the fluid properties and energy requirements of the system for several situations. It is assumed that the condenser and turbine operate at the same pressure and temperature. These situations are as follows: Case 1: cooling load = 5 kW, electrical load = 1 kW, position = 20°N lat. and ambient temperature = 28°C. To be powered by solar energy alone. Case 2: cooling load = 40 kW, electrical load = 2 kW, position = 50°N lat. and ambient temperature = 18°C. To be powered equally by gas and solar energy. Case 3: cooling l o a d = 10kW, electrical l o a d = 10kW, p o s i t i o n = 30°N lat. and ambient temperature = 18°C. To be powered equally by gas and solar energy. The collector efficiency was obtained from data given by Stoecker and Jones [6]. A double-glazed non-selective absorber was used in each case. The open-cycle salt-water desalination system was analysed for the lithium-bromide-water working pair over a range of different operating conditions. It was assumed that the vapour pressure for sea water was the same as that for pure water at the same temperature. In fact, the vapour pressure of 34% saturated brine (typical sea water) is only 2% below that of pure water. It was also assumed that the heat-exchanger efficiency was 80%, as in the previous example. In the proceeding text the performance of the cycle is expressed as a percentage and is the minimum theoretical energy needed (3.49 kJ kg- ~) to separate the water from the salt in 'standard'

352

M. Nguyen et al. (a)

0.80 EVAPORATOR TEMR= 5°C ABSORBER TEMR= 20°C

0.75 0.70 Q. 0.65 O O 0.60 0.55 0.50 0.45

70

80

I

I

I

I

I

I

I

T

I

I

90

100

110

120

130

140

150

160

170

180

GENERATOR TEMPERATURE ( ~ )

(b)

25 C 2O >o z

TC= 40"C

,,u. 15 w w _z m ¢Y

Tc = 30 *C

J EVAPORATOR TEMP.= 5°C ABSORBER

TEMP.=

20°C

T(; = 20"C

70

I

I

80

90

I

I

I

i

I

l

100 110 120 130 140 150 GENERATOR TEMPERATURE (°(3)

I

I

I

160

170

180

Fig. 4. (a) Variation of COP with generator and condenser temperature. (b) Variation of turbine efficiency with generator and condenser temperature.

brine, divided by the actual energy required by this system. This may be likened to an efficiency and the performance described in this manner is more easily compared to that of other systems. Figure 7 shows the effect of evaporator temperature on the performance of the open-cycle water desalination system for a range of generator temperatures. It can be seen that the efficiency of the desalination cycle increases with increasing evaporator temperature. Unlike the combined refrigeration and power-generation system, the temperature of the evaporator in the water desalination system is not constrained by cooling requirements. In the above analysis, the maximum evaporator temperature is constrained to be somewhat below the absorber temperature. As condenser temperature decreases, the efficiency of the desalination system increases. Ideally, for maximum performance the condenser should be operated at as low a temperature as possible (Fig. 8), however, this temperature will be constrained by that of the condenser coolant, i.e. ambient air temperature.

Solar/gas-driven a b s o r p t i o n h e a t - p u m p systems

353

Figure 9 shows the effect of absorber temperature on the performance of the system for a range of generator temperatures. It can be seen that the efficiency of the system decreases with increasing absorber temperature. Ideally, an absorber temperature as low as possible should be used, the lowest temperature being constrained by the absorber coolant temperature. If the absorber is air-cooled then this minimum temperature will be slightly above ambient. The temperatures of the absorber and the evaporator of a working system must be chosen carefully, bearing in mind that high evaporator temperatures are desirable but the maximum temperature is constrained by the absorber temperature. Calculations show that under favourable conditions this system can produce 8 litres of water per square metre of solar collector area per day in equatorial regions without any contribution of heating from gas. This figure is an average over the summer months. Typically solar stills average 4 litres of water per square metre per day. Whilst the figure given for this absorption desalination system is theoretical and some simplifications have been made, its productivity is still likely to be somewhat greater in practice than that of the solar still. The use of gas in combination with solar energy in the above system would further extend its productivity.

(a)

0.80 I ~ ITE = 1 0 O C ~ , , ~

0.40 / I 70

(b)

CONDENSER TEMP.= 45°C ABSORBER TEMP.= 20°C

I TE =1"10°C I I I I I I 80 90 100 110 120 130 140 150 GENERATOR TEMPERATURE (°C)

r 160

I 170

I 180

25 TE =0*C

j

20 o u. m

=_zl o m

rE = -s°~/

IZ

5

CONDENSER TEMP.= 45 °C ABSORBER TEMP.= 20 °(3

T~ = -10*C

0

70

i

[

80

90

i

i

i

i

i

i

100 110 120 130 140 150 GENERATOR TEMPERATURE (°C)

i

i

i

180

170

180

Fig. 5. (a) Variation of C O P with generator and e v a p o r a t o r temperature. (b) Variation of turbine efficiency with generator and e v a p o r a t o r temperature.

354

M. Nguyen et al. 080 ~ _

EVAPORATOR TEMP.= 5° C CONDENSER TEMP.= 45OC

~

0.70

0.60 ~ 0.50 0 0.40

0.30 0.20

~ 90

80

(b)

25

r 100

I ~ I ~ t 110 120 130 140 150 GENERATOR TEMPERATURE (°(3)

r 160

I 170

r 180

r T. = 200C

~20

0.,oc 80

90

7 100

110 120 130 140 150 GENERATOR TEMPERATURE (°C)

160

170

180

Fig. 6. (a) Variation of COP with generator and absorber temperature. (b) Variation of turbine efficiency with generator and absorber temperature.

0.096

0.092

0.088 E

Inlet water temp.= 15 °C Absorber temp.= 30 "13 Condenser temp.= 40 *C

LU

0.084

"~... Region of crystallisation

0.08 50

, 60

, 70

t 80

, 90

J 100

, 110

, 12(

Generator temperature ( * C ) Fig. 7. Effect of generator and evaporator temperature on desalination efficiency.

Solar/gas-driven absorption heat-pump systems

355

Table 1. Case 1

Case 2

120 70 10 30 28 613 3289 329 1273 --84 --84 139 301 --32

70 40 10 20 18 613 1553 179 1273 --76 --76 66 81 -- 161

- 32

--161

To (°C) Tc (°C) TE CC) T, (°C) r~ PE (kPa) Pc (kPa) h I (kJ kg- i ) h; (kJ k g - i h 3 (kJ kg 1 ha (kJ kg i h 5 (kJ k g - i h6 (kJ k g - 1 h7 (kJ kg h s (kJ kg ~) h 9 (kJ kg ~) hlo (kJ k g - 1 ) m 3 (kgs -I ) m 6 (kgs -l ) m 9 (kg s ~) m,o (kg s =) COP nr (%) (2o (kW) Solar insolation (W m - 2 ) Collector efficiency (%) Collector area (m 2) Gas flow rate (kgs i)

1554 1554 0.0268 0.0179 0.00530 0.00356 0.553 16.1 15.4 1020 52 29 --

Case 3 120 70 10 20 188 613 3289 329 1273 --76 -- 76 95 301 --77 --77 1554 1554 0.0860 0.0398 0.0106 0.0356 0.577 17.2 75.6 980 60 64 0.00114

1373 1373 0.138 0.813 0.0366 0.0198 0.823 7.6 74.9 790 68 70 0.00112

Subscripts refer to the positions labelled in Fig. 2.

0.098 .re. 2s=c

Inletwater temp.= 15 *C E~porator temp.= 20 °C Absorbertemp.= 30 *C

0.096 .-~c

0.094

t-,,

~~ : : : : . r¢-~c T¢"~¢"~

"---.. Regionofc~/stalllsetion :::iii~!-......

"

~= 0 . 0 9 2 HJ

5 'o

0,088 40

60'

;o

80'

100 '

9 'o

110 '

120 '

Generator temperature (°C) Fig. 8. Effect o f g e n e r a t o r a n d c o n d e n s e r t e m p e r a t u r e o n d e s a l i n a t i o n efficiency.

0.098 [ /

["

o.o. t /

Inletwater temp.= 15 °C

~

Evaporetortemp.=12 °C

oo _.ooo=

[- /

~"

°~ 0.090 IT" ,o.o

T,.,o~----..~

WUJ

-i i ............... .......

T~-~°c

40

50

60

70

.

"".....Region

of crystallisetion

..............iii~i~...........................

..............

" ...............

80

90

"....................

100

110

120

G e n e r a t o r temperature ( ° C ) Fig. 9. Effect o f g e n e r a t o r a n d a b s o r b e r t e m p e r a t u r e o n d e s a l i n a t i o n efficiency.

356

M. Nguyen et al. 4. C O N C L U S I O N S

T h e r e is c o n s i d e r a b l e scope for the a p p l i c a t i o n o f h y b r i d g a s / s o l a r - a b s o r p t i o n systems where i n t e r m i t t e n t o r low solar i n s o l a t i o n c u r r e n t l y restricts their use. The w o r k i n g fluids a n d the energy sources are a t t r a c t i v e on e n v i r o n m e n t a l grounds. C a l c u l a t i o n s for a c o m b i n e d refrigeration a n d p o w e r - g e n e r a t i o n system o p e r a t i n g on g a s / s o l a r energy s h o w t h a t a C O P a o f 0.75 a n d a r/T o f 16% are achievable simultaneously. F u r t h e r increases in p e r f o r m a n c e m a y be achievable using multi-effect systems, A s o l a r - p o w e r e d w a t e r d e s a l i n a t i o n system c a n achieve a yield o f 8 litres p e r square m e t r e o f collector a r e a p e r d a y , well b e y o n d that o f a solar still. T h e use o f gas in c o m b i n a t i o n with solar energy will increase this yield further. F u r t h e r research into the e c o n o m i c a l viability o f such systems is required to assess their c o m m e r c i a l potential. Such research s h o u l d t a k e into a c c o u n t a n y a n t i c i p a t e d changes in legislation g o v e r n i n g the use a n d m a n u f a c t u r e o f c o n v e n t i o n a l refrigerants. Acknowledgement--The authors would like to thank British Gas for its financial and technical support of this project.

REFERENCES 1. B. M. Cohen and D. Kosar, Solar Water Heating with Natural Gas Backup. Gas Research Institute, Chicago, Illinois, U.S.A. 2. S. B. Riffat and N. J. Shankland, Comparison of R134a and R12 refrigerants in vapour compression systems. Int. J. Energy Res. 17, 439-442 (1993). 3. D. J. G. Butler, Phase-out of CFCs and HCFCs: options for replacement and containment. CIBSE National Conf. (1994). 4. T. N. Veziroglu and F. Barbir, Initiation of hydrogen energy systems in developing countries. Int. J. Hydrogen Energy 17, 527-538 (1992). 5. K. S. Spiegler, Salt Water Purification. John Wiley & Sons, New York (1962). 6. W. F. Stoecker and J. W. Jones, Refrigeration and Air-conditioning, 2nd Edn, p. 389. McGraw-Hill, New York (1982).