Simulation and Analysis of Biogas operated Double Effect GAX Absorption Refrigeration System

Transcription

1 Simulation and Analysis of Biogas operated Double Effect GAX Absorption Refrigeration System Simulation and Analysis of Biogas operated Double Effect GAX Absorption Refrigeration System G Subba Rao, 2 Vemuri Laksminarayana Dept of Mecanical Engineering, 2 Dept of Mecanical Engineering Geetanjali Institute of Science and Tecnology, Gangavaram (V), Kovur (M), NelloreDt, , Andra Prades, India, 2 Velammal Institute of Tecnology, Cennai, Tamilnadu, India subbrow@gmailcom, vlnvamsi@yaoocom Abstract : A termodynamic simulation of a double effect generator eat ecanger absorption refrigeration cycle using biogas as source of energy as been carried out Te binary miture considered in te present investigation was NH 3 H 2O (Ammonia - Water) Tis simulation was performed in order to investigate te effect of te temperature and pressure of te ig temperature generator and te pressure of evaporator ave over te Coefficient of Performance (COP) for a constant condenser and absorber temperatures Te basic parameters at various state points of te cycle was computed using standard correlations Te solution circulation rates and volume of biogas required for operation of te cycle are analysed for te variations in operating parameters at te ig temperature generator and evaporator Keywords: Absorption refrigeration, Double effect cycle, Binary miture, Coefficient of performance, Solution circulation rate, Termodynamic Simulation I INTRODUCTION Te need to augment world s energy sources as recently stimulated tose interested in renewable energy sources and waste eat utilization to reeamine te potential of absorption cycles for refrigeration applications Toug, te absorption cycles ave low COP values, wit increasing energy prices, Ozone layer depletion, Global warming effects, tese units are going to become more competitive in te days to come Unlike mecanical vapor compression refrigerators, tese systems cause no ozone depletion and reduce demand on electricity supply Besides, eat powered systems could be superior to electricity powered systems because of te use of inepensive waste eat, solar, biogas, biomass or geotermal energy sources for wic te supplying cost is negligible in many cases Despite using an economic energy source, te system is caracterized by its low COP, for tat reason it is necessary to perform a study in order to find te most efficient operation range One of te main factors tat ave elped to develop tis kind of systems is te termodynamic simulation tat can be carried out in order to study te different variables affecting te performance of te equipment In 77 Swarts and Sitzer analyzed termodynamically te possibility to operate te solar absorption refrigeration system for air conditioning Teir results sowed tat te system was suitable for domestic use Van Passen 2 presented te termodynamic simulation of a solar absorption refrigeration system Witlow 3 studied te absorption refrigeration cycle from te termodynamic point of view Te use of eat ecangers and some oter binary mitures were recommended Da Wen Sun 4, analyzed and performed an optimization of te water ammonia cycle As a result, e obtained a matematical model tat allowed te simulation of te process Sun 5, presented a termodynamic design and performed an optimization of te absorption refrigeration process in order to map te most common cycles for water ammonia, and litium bromide water Te results can be used to select te operation conditions in order to obtain a maimum performance from te system Sun 6, performed a termodynamic analysis of different binary mitures considered in te absorption refrigeration cycle A lot of work as been done in tis area and te effect of te generator and evaporator temperatures ave been considered etensively MA Siddiqui et al 5 as studied te optimum generator temperatures for single effect ammonia water absorption systems at subfreezing evaporator temperatures In tis work, a double effect vapour absorption refrigeration system operating on NH 3 H 2 O as refrigerant and absorbent pair is considered for analysis Initially, te basic properties at various state points of te cycle was computed using standard correlations Te ig temperature generator temperature and pressure and te evaporator pressure were varied and te effect of te variation over te Coefficient of Performance (COP) for a constant condenser and absorber temperatures ave been studied Te solution circulation rates and volume of International Journal of Applied Researc in Mecanical Engineering, Volume-, Issue-, 20

2 biogas required for operation of te cycle are analysed for te variations in operating parameters at te ig temperature generator and at te evaporator II SYSTEM DESCRIPTION : DOUBLE EFFECT VAPOUR ABSORPTION REFRIGERATION SYSTEMS A number of different configurations can produce a double effect absorption ciller Te two basic types are te double-condenser double effect and te double-absorber double effect Teir principle is based on te fact tat te cooling capacity depends primarily on te amount of refrigerant tat is vaporized in te evaporator and tat by reusing te waste eat from te condensation or te absorption stages, more refrigerant can be desorbed from solution In te former case, a primary ig temperature generator yields vapor wose latent eat is used to fire a secondary low temperature generator at a lower pressure Te latent eat of te vapor desorbed in te secondary generator is ten rejected to te eat sink In te latter case eat from one absorber is used to fire te secondary generator Te vapor from bot generators is combined and enters a single condenser Fig illustrates te main components of te double effect absorption refrigeration cycle Hig-pressure liquid refrigerant from te condenser passes into te evaporator troug a precooler and an epansion valve (V-4) tat reduces te pressure of te refrigerant to te low pressure eisting in te evaporator Te liquid refrigerant vaporizes in te evaporator by absorbing eat from te material being cooled and te resulting low-pressure vapor passes to te absorber troug a precooler, were it is absorbed by te strong solution coming from te low temperature generator troug an epansion valve (V-), and forms te weak solution Te weak solution is pumped to te ig temperature generator pressure troug Preeater I and Preeater - II, and te refrigerant is boiled off Te remaining solution flows first to te low temperature generator troug Preeater II Te refrigerant boiled off from te ig temperature generator is passed troug te low temperature generator inside a pipe were eat ecange takes place tereby furter liberating refrigerant from te solution in te low temperature generator Te refrigerant generated from te low temperature generator and te ig temperature generator enter te condenser Te weak solution from te low temperature generator is passed troug te Preeater I to te absorber were it absorbs te refrigerant vapours coming from te evaporator tus completing te double effect cycle By weak solution is meant tat te ability of te solution to absorb te refrigerant vapor is weak, according to te ASHRAE definition In order to improve system performance, te preeaters are included in te cycle An analyzer and a rectifier need to be added to remove water vapor from te refrigerant miture leaving te generator before reacing te condenser For te current study, it is assumed tat te refrigerant vapor is 00% ammonia and te analysis of analyser and rectifier ave not been considered in tis work III SYSTEM MODELING AND SIMULATION In order to analyze te system, mass and energy balance must be performed at eac component Te mass, material and energy balance at eac component as a control volume gives te following relations At absorber 0 m = m + m () m = m0 0 + (2) Q A m = m + m00 m (3) At ig temperature generator 4 5 m = m + m (4) m 4 4 = m5 5 + (5) Q G m = m + m55 m44 (6) At low temperature generator 7 8 m = m + m (7) m 7 7 = m8 8 + (8) At condenser Q C = m33 (9) At evaporator Q E = m (0) + m m m m International Journal of Applied Researc in Mecanical Engineering, Volume-, Issue-, 20

3 Pre - Cooler 3 2 Condenser V-3 Low Temperature 7 V-2 Desorber V V- 9 Hig Temperature Desorber Pre - Heater- II Pre - Heater- I Evaporator 2 Absorber Solution Pump Fig: Double Effect GAX Vapour Absorption Refrigeration System General m = m = m = m () m = m = m (2) 9 0 m = m = m (3) 2 3 m = m = m () m = m = m = m = m = m () () () = () It is assumed tat pure Ammonia is entering te condenser ie, (=) yields 2 3 = 0 () COP Q Q = E/ G (20) Work input to solution pump is computed and included in te calculations Termodynamic Properties Equilibrium pressure data of pure Ammonia, Liquid entalpy for NH 3 - H 2 O miture are from Zeigler & Trepp Liquid and vapour entalpy for NH 3, H 2 O and NH 3 - H 2 O miture are from Infante Ferriera Equilibrium pressure data of NH 3, H 2 O miture is from Perry and Clinton Te correlations for Supereated vapour entalpy for NH 3 miture was taken from CPArora Te eating values of te biogas and te volume of te biogas required per TR of refrigeration were calculated using te relations developed by MA Siddiqui et al In tis study a computer code as been developed to compute te first law analysis of te absorption ciller A detailed analysis of te absorption ciller requires a knowledge of main system pressures wic are maimum, intermediate and minimum pressures, temperatures and flow rates at strategic points in te International Journal of Applied Researc in Mecanical Engineering, Volume-, Issue-, 20

4 system In te simulation te following operating parameters i) HT generator temperature ii) HT generator pressure and iii) Evaporator pressure were varied and te effect of te variations are studied to arrive at te optimum operating conditions A number of model runs ave been performed and compared in order to investigate te interactions of different operating conditions on te performance of te absorption unit Te following assumptions were made during te analysis: Te condenser temperature is kept equal to te absorber temperature 2 Heat losses and gains between te system and its environment are neglected 3 Friction and pressure losses in pipes and components are neglected For te design values assumed initially Absorber = 25 0 C, Hig temp Generator = 0 0 C, Evaporator = -0 0 C, Condenser = 25 0 C, t B = biogas source temperature at HTG inlet = C; Capacity = kw; Te component analysis is done as follows: Evaporator : Assuming, t 20 = 6 0 C, t 2 = 2 0 C, Q E-I = m * ( ) = 0 kw; ΔT E-I =[(t 2 t E-I ) (t 20 t E-I )] / ln [(t 2 t E-I ) / (t 20 t E-I ) ] ΔT E-I = ; Evaporator Area = A E-I = Q E-I / K E-I ΔT E-I = m 2 ; Preeater I : Q PH-I = m 2* ( 3 2 ) = kw; ΔT LTHE-I =[(t 8 t 3 ) (t 9 t 2 )] / ln [(t 8 t 3 ) / (t 9 t 2 ) ] = ; Preeater I area = A PH-I = Q PH-I / K PH-I ΔT PH-I = m 2 ; Preeater II : Q PH-II = m 3* ( 4 3 ) = kw; ΔT PH-II = [(t 5 t 4 ) (t 6 t 3 )] / ln [(t 5 t 4 ) / (t 6 t 3 )] = ; A PH-II = Q PH-II / K PH-II ΔT PH-II = m 2 ; Precooler : Q PC = m * ( ) = kw; ΔT PC = [(t t ) (t t )] / ln [(t t ) / (t t )] = ; A PC = Q PC / K PC ΔT PC = m 2 ; IV RESULTS AND DISCUSSION Figs 2 & 3 sow te variation of COP against te HT generator pressures for different temperatures keeping te evaporator pressure constant at 20 kpa, condenser and absorber temperatures constant at 40 0 C Marked improvement in COP is observed wen te system is operated at generator pressures between 2800 and 4000 kpa It is also observed tat moderate HT generator pressures and lower temperatures sow better operating performance and yields a iger COP It is also observed tat at iger HT generator pressures require iger temperatures and tus te COP curves become almost flat as it tends to increase te average temperatures in te condenser and absorber Hig temperature Generator : Q HTG = m * + m 5* 5 m 4* 4 = kw; ΔT HTG = [(t B t HTG ) (t B t )] / ln [(t B t HTG ) / (t B t )] ΔT HTG = ; Hig temperature Genarator Area = A HTG = Q HTG / K HTG* ΔT HTG = m 2 ; COP Condenser : Assuming t 22 = 34 0 C, t 23 = 38 0 C, Q C = m * + m 3* 3 m * = kw; ΔT C = [(t C t 22 ) (t C t 23 )] / ln [(t C t 22 ) / (t C t 23 )] = ; Condenser area = A C = Q C / K C ΔT C = m 2 ; Absorber : Assuming, t 24 = 32 0 C, t 25 = 38 0 C, Q A = m * + m 0* 0 m * = kw; ΔT A = [(t A t 24 ) (t A t 25 )] / ln [(t A t 24 ) / (t A t 25 )] = ; Absorber area = A A = Q A / K A ΔT A = m 2 ; Pgen (kpa) T tg = 0 C T tg = 0 C T tg = 0 C T tg = 0 C T tg = 0 C T tg = 200 C Fig 2 COP vs variation of pressure at ig temperature Generator International Journal of Applied Researc in Mecanical Engineering, Volume-, Issue-, 20

5 0 COP T t gen (C) P tg = 2000 kpa P tg = 2200 kpa P tg = 2400 kpa P tg = 2600 kpa P tg = 2800 kpa P tg = 3000 kpa Ptg = 3200 kpa Soln circ rate (kg/s) P t gen (lpa) T tg = 0 C T tg = 0 C T tg = 0 C T tg = 0 C T tg = 0 C T tg = 200 C Fig 3 COP vs variation of temperature at ig temperature Generator Fig 4 sows te minimum volume of biogas required for running te cycle for variations in HT generator pressure Te requirement of biogas reduces as te generator pressure increases due to te fact tat as te COP increases te quantity of eat input required decreases accordingly It can also be noted tat at low HT generator temperatures requirement of biogas is low indicating generation of low temperatures at te desorber need reduced energy inputs Fig 5 Variation of pressure at ig temperature Generator vs Solution Circulation rate COP Vol of biogas (cu m / day) P evep (kpa) T t gen = 0 C T t gen = 0 C T t gen = 0 C T t grn = 0 C T t gen = 0 C Fig 6 COP vs variation of pressure at evaporator P t gen (kpa) T tg = 0 C T tg = 0 C T tg = 0 C T tg = 0 C T tg = 0 C T tg = 200 C Fig 4 Variation of pressure at ig temperature Generator vs Vol of Biogas requirement Fig 5 indicates increased solution circulation rates for increase in HT generator pressure and it decreases wit te increase in HT generator temperatures Wit te increase in solution circulation rates eiter te cooling capacity increases or te energy input required for te same cooling capacity reduces Fig 6 gives information about te pressure variations at te evaporator Te COP reduces wit te increase in evaporator pressure drastically and tereafter remains almost constant for any furter increase V CONCLUSION Ammonia water absorption refrigeration cycle was analyzed, wit te termodynamic properties at strategic points being calculated from te establised correlations Te coefficient of performance (COP) of tis cycle versus ig temperature generator pressure and temperature and evaporator pressure was analyzed and it was noticed tat tese parameters are important in determining te optimum operating conditions for te system Moderate HT generator pressures and lower temperatures yield good results and better performance of te system Similarly lower HT generator temperatures result in reduced requirements of energy inputs and ence low quantities of biogas is sufficient to power te absorption cycle Better performance of te system is observed wen te generator temperatures are kept low for any fied evaporator pressure Wit te variations in pressures and temperatures at te HT generator and evaporator te solution circulation rates are greatly influenced International Journal of Applied Researc in Mecanical Engineering, Volume-, Issue-, 20

6 REFERENCES [] I Swartz, and A Sitzer, Solar Absorption System for Space Cooling & Heating, ASHRAE Journal,, (), 77, 5-54 [2] J P Van Passen, Solar Powered Refrigeration by means of an Ammonia-Water Intermittent Absorption Cycle Ed (May 87); p2 [3] EP Witlow, Tends of Efficiencies in Absorption Macines, ASHRAE Journal,, (), 66, 44 [4] DW Sun, Computer Simulation and Optimization of Ammonia-Water Absorption Refrigeration Systems, Energy Sources,, (3), 97, 2-22 [5] DW Sun, Termodynamic Design Data an Optimum Design Maps for Absorption Refrigeration Systems, Applied Termal Engineering,, (3), 96, 2-22 [6] DW Sun, Comparison of te Performances of NH3- H2O, NH3-LiNO3 and NH3-NaSCN Absorption Refrigeration Systems: Energy Conversion Management, 39, (5/6), 98, [7] ASHRAE, ASHRAE Handbook, Refrigeration Systems and Applications, Capter 40, p 40, ASHRAE, 9 Tullie Circle, N E, Atlanta, GA 30329, 94 [8] Ataer, OE, Gogus, Y 9 Comparative study of irreversibilities in aqua-ammonia absorption refrigeration systems International Journal of Refrigeration : [9] Izquierdo, M, Aroca, S 90 Litium bromide ig temperature absorption eat pump: Coefficient of performance and eergetic efficiency International Journal of Refrigeration : [0] Koeler, WJ, Ibele, WE, Soltes, J, Winter, ER 88 Availability simulation of a litium bromide absorption eat pump Heat Recovery System & CHP 8(2): 7- [] Karakas, A, Egrican, N, N, Uygur, S 90 Second law analysis of solar absorption cooling cycles using litium/water and ammonia/water as working fluids Applied Energy 37: 7-7 [2] Sun-Fu Lee, Serif, SA 99 Second-law analysis of multi effect litium bromide/water absorption cillers ASHRAE Transactions 23(3): [3] Sun-Fu Lee, Serif, SA 200 Second-law analysis of double effect litium bromide/water absorption cillers ASHRAE Transactions 9(5): [] Subba Rao, G, Vemuri Laksminarayana Eperimental investigation of an industrial double effect LiBr - H 2O Vapour absorption refrigeration system 2005 Proc of national conference on Emerging Trends in Energy and Environment, Dept of Environment, Tamilnadu & Sairam Engineering College, Cennai, India pp: [] Apornratana S, and I W Eames, 95 Termodynamic analysis of absorption refrigeration cycles using second law of termodynamic metod International Journal of Refrigeration (4): [] ASHRAE ASHRAE Handbook- Fundamentals Atlanta: American Society of Heating, Refrigeration and Air-Conditioning Engineers, Inc [] Bejan A, G Tsatsaronis, and M Moran 96 Termal design and optimization New York: Jon Wiley [] Eisa M A R, and F A Holland 86 A study of te operating parameters in a water litium bromide absorption cooler International Journal of Energy Reasearc 0: 37-4 [] Grossman G, MWilk and R C DeVault 94 Simulation and performance analysis of triple-effect absorption cycles ASHRAE Transactions 00(): [20] Van Wylen G J, Sonntag R E 76 Fundamentals of classic termodynamics 2 nd edn Jon Wiley & Sons, New York International Journal of Applied Researc in Mecanical Engineering, Volume-, Issue-, 20