researcharticle analysis of hybrid ejector absorption
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Research ArticleAnalysis of Hybrid Ejector Absorption Cooling System
Doniazed Sioud and Ahmed Bellagi
Department of Energy Engineering, Ecole Nationale dβIngenieurs de Monastir (ENIM), University of Monastir, Tunisia
Correspondence should be addressed to Doniazed Sioud; siouddoniazed@gmail.com
Received 17 July 2018; Revised 25 February 2019; Accepted 17 March 2019; Published 2 September 2019
Academic Editor: Oronzio Manca
Copyright Β© 2019 Doniazed Sioud and Ahmed Bellagi. This is an open access article distributed under the Creative CommonsAttribution License, which permits unrestricted use, distribution, and reproduction in any medium, provided the original work isproperly cited.
In this paper, a hybrid ejector single-effect lithium-bromide water cycle is theoretically investigated. The system is a conventionalsingle-effect cycle activated by an external steam-ejector loop. Amathematicalmodel of the whole system is developed. Simulationsare carried out to study the effect of the major parameters of the hybrid cycle on its performances and in comparison with theconventional cycle. The ejector performance is also investigated. Results show that the entrainment ratio rises with steam pressureand condenser temperature, while it decreases with increasing generator temperature.The effect of the evaporator temperature onejector performance is negligible. It is shown also that the hybrid cycle exhibits better performances than the corresponding basiccycle. However, the performance improvement is limited to a specific range of the operating parameters. Outside this range, thehybrid system behaves similar to a conventional cycle. Inside this range, theπΆππ increases, reaches amaximum, and then decreasesand rejoins the behavior of the basic cycle.The maximum πΆππ, which can be as large as that of a conventional double-effect cycle,about 1, is obtained at lower temperatures than in the case of single-effect cycles.
1. Introduction
Cooling and air conditioning are essential for small scale andlarge industrial process applications. While systems applyingthe vapor-compression technique use environmental harmfulrefrigerants (FCC, FCHC, etc.), absorption technique forproduction of cold is based on environment friendly workingfluids, namely, aqueous lithium bromide solutions with wateras refrigerant or water-ammonia mixtures with ammonia asrefrigerant. This technique however suffers from low perfor-mances. That is the reason why new hybrid and combinedconfigurations are proposed, implying the integration of newcomponents, particularly ejectors, in order to enhance theperformances.
Various configurations incorporating ejectors were stud-ied. Exhaustive review of the literature on this subject canbe found in Besagni et al. [1, 2]. Elaborated CFD-models ofejectors developed to evaluate the ejector performances inboth on-design and off-design conditions have been also pub-lished [3]. Combined cycles were investigated with ejector setat the absorber inlet [4β9]. πΆππ of such cycles are reportedto be higher by about 2β4% than that of conventional cycles.Principally, investigations indicate that πΆππ of the combined
configuration are greater or equal to that of single-effectcycles, but reached at lower generator temperatures.
Other configurations are discussed where the ejector islocated at the condenser inlet of single-effect systems [10β14].Theoretical investigations confirm the improvement of theperformances in comparison with basic single-effect cycles.Experimental studies [15] show that this combined cycle is 30-60% more performant than conventional absorption cyclesand almost reaches theπΆππ of double-effect systems. Besidesmodifying configurations, adding a flash tank between ejec-tor and evaporator was also proposed [16, 17].
Ejector improved double-effect absorption system wasalso investigated [18β20].TheπΆππ of the proposed refrigera-tion schemewas found to increasewith the temperature of theheat source until this temperature reaches 150βC. Beyond thatvalue, the new cycle worked as a conventional double-effectcycle. Another configuration was studied with an ejector cou-pled to vapor generator [21β23]. This procedure is intendedto enhance the concentration process by compressing thevapor produced from the lithium bromide solution in orderto reheat the solution fromwhich it came. Results showed thatπΆππ of the new cycle increases especially with the heat sourcetemperature.
HindawiJournal of EngineeringVolume 2019, Article ID 1862917, 13 pageshttps://doi.org/10.1155/2019/1862917
2 Journal of Engineering
4
5
611
QG
QAB
3
7
1
2
9
10
Condenser
EvaporatorAbsorber
Generator Q CD
Q EV
(a)
Evaporator
Generator
Absorber
SteamGenerator
Condenser
Q EVQAB
QCD
QSG
11
8
16
1719
18
7
9
10
13 14
4
5
6
3
1
2
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12
(b)
Figure 1: Single-effect absorption system: (a) conventional; (b) hybrid, ejector-enhanced.
In this paper, an ejector-activated single-effect LiBr-water cycle is proposed and theoretically investigated. Theobjective is to assess the feasibility and limits of performanceof this new cycle scheme. If the πΆππ of the proposedsystem could reach that of a conventional double-effect cycle,this would mean obtaining high performance by avoidingthe configuration complexity of double-effect cycles. Weinvestigate the evolution of the πΆππ of the hybrid cycle withthe steam generator temperature and the main factors of thecooling machine, i.e., desorber, condenser, and evaporatortemperature.The behavior of the entrainment ratio as ejectorperformance criterion is also investigated for various primaryand secondary flow pressure and backpressure.
2. System Description
Figures 1(a) and 1(b) are schematics of a conventional single-effect absorption cycle and an ejector-enhanced single-effectabsorption system. A conventional single-effect absorptionchiller (Figure 1(a)) is composed of evaporator, absorber,condenser, generator, solution expansion-valve, pump, solu-tion heat exchanger, and refrigerant expansion-valve. In ahybrid system (Figure 1(b)) a steam-generator-ejector loopis coupled to the conventional single-effect installation viathe machine generator. This extra circuit is constituted of anejector, a steam generator, a water pump, and an expansionvalve.
Journal of Engineering 3
Primaryfluid
Secondaryfluid
CONSTANT AREASECTION
NOZZLESECTION
DIFFUSER SECTION
Back
-pre
ssur
e
Noz
zle ex
it pl
ane (
i)
Plan
e (j)
Plan
e (k)
18
19
12At
Ai Ac
Figure 2: Ejector schematics.
The ejector loop is intended to improve the cycle per-formance by enhancing the concentration process in themachine generator. A high-pressure flow (18) coming fromthe external steam generator enters the primary nozzle ofthe ejector where its pressure drops while it is accelerated.At the nozzle exit section (π) (Figure 2) its velocity becomessupersonic and high enough to entrain a secondary flow (19in Figure 1), part of the vapor (7) generated in the desorber.The two streams mix in the mixing chamber and the resultinggas, after undergoing a shockwave that reduces its velocityto subsonic, is compressed in the diffuser forming the lastsegment of the ejector. The exiting vapor (12) condenses inthe coil placed inside the solution generator, liberating thuscondensation heat used to concentrate the saline solution bydesorbing vapor from the water-rich solution (3) enteringthe generator. Part of the condensate flows, after appropriatepressure reduction, to the condenser, and the rest is pumpedback to the steam generator.
3. Chiller Model
Basing on mass and energy balances written for everymachine element a mathematical model of the installationis set up. For the numerical simulations, a computer code ofthemachinemodel is realized using the software EngineeringEquations Solver, EES [24].
Themodel is elaborated under the following assumptions:
(i) Steady state conditions
(ii) Negligible heat losses to the surroundings at genera-tor, condenser, absorber, and evaporator
(iii) Negligible pressure losses in pipes and components
(iv) Saturated refrigerant exiting condenser and evapora-tor
(v) Isenthalpic flow in solution and refrigerant valves
(vi) Phase equilibrium between solution entering refrig-erant generator and vapor leaving
(vii) Constant solution flow-rate leaving the absorber,specifically 2 kg/s
(viii) Heat exchanger effectiveness, πHX = 80%In the following major elements of the model are presented.
3.1. Ejector Loop. This loop includes steam generator, ejector,heating coil placed in solution generator, expansion valve,and water pump.
(i) Steam Generator
The mass and energy balances on steam generator write,respectively,
οΏ½οΏ½17 = οΏ½οΏ½18 (1)
οΏ½οΏ½ππΊ = οΏ½οΏ½17 (β18 β β17) (2)
The properties of exiting saturated vapor (18) are:πππΊ = π18 = ππβπ ππ‘ (π18) (3)
β18 = βπβπ ππ‘ (π18, π18 = 1) (4)
Further, π17 = π18 .Properties with index π for water refer to pure water
properties as given in steam tables.
(ii) Ejector
The ejector performance depends on the backpressureπππβthe pressure of the exiting (supposed saturated) steamflowing in the heating coilβ, the primary pressure, π18 ,and the secondary pressure, π19. The relations between thedifferent pressures around the ejector are
πππ = π13 = ππβπ ππ‘ (π13) (5)
π19 = π7 = π8 (6)
The mass balance for the ejector writes
οΏ½οΏ½12 = οΏ½οΏ½18 + οΏ½οΏ½19 = (1 + π) οΏ½οΏ½18 (7)
4 Journal of Engineering
where π stands for the entrainment ratio
π = οΏ½οΏ½19οΏ½οΏ½18
(8)
The enthalpy of exiting flow (12) can be deduced from theenergy balance
β12 = β18 + πβ191 + π (9)
(iii) Heating Coil
Assuming a difference of 5 K between the temperatures of theheat source and that of the refrigerant generator solution, weget
π13 = π12 = ππΊ + 5 = π4 + 5 (10)
β13 = βπβπ ππ‘ (π13, π13 = 0) (11)
The mass balance writes
οΏ½οΏ½12 = οΏ½οΏ½13 (12)
(iv) Water Pump
We suppose approximately isothermal pumping
π17 = π16 = π13 = π14 (13)
The mass and energy balances write, successively,
οΏ½οΏ½17 = οΏ½οΏ½16 (14)
β17 = β16 + (π17 β π16)π17 (15)
where the term [(π17βπ16)/π17] in the last equation representsthe specific pump work (kJ/kg), with (π17 = ππ(π17, π17)).
(v) Expansion Valve
The expansion is isenthalpic, i.e.,
β14 = β15 = β13 (16)
οΏ½οΏ½14 = οΏ½οΏ½15 (17)
3.2. Liquid Solution Loop. The absorber-generator loop com-prises absorber, solution valve, solution pump, solution heatexchanger, and refrigerant generator.
(i) Refrigerant Generator
With π denoting the lithium bromide concentration in theliquid solution, the mass balances for this machine elementwrite
οΏ½οΏ½3 = οΏ½οΏ½7 + οΏ½οΏ½4 (18)
οΏ½οΏ½4π4 = οΏ½οΏ½3π3 (19)
Solving for οΏ½οΏ½7 yields
οΏ½οΏ½7 = οΏ½οΏ½4
π4 β π3π3 (20)
For the energy balance we get
οΏ½οΏ½4β4 + οΏ½οΏ½7β7 = οΏ½οΏ½3β3 + οΏ½οΏ½12 (β12 β β13) (21)
from which we deduce
οΏ½οΏ½4 = οΏ½οΏ½12 (β12 β β13) β (οΏ½οΏ½7β7 β οΏ½οΏ½3β3)β4
(22)
The properties of water-weak solution (4) exiting the genera-tor are determined as follows:
ππΊ = ππΆπ· = π4 (23)
π4 = ππππΏβπ ππ‘ (ππΊ, ππΊ) (24)
β4 = βπππΏβπ ππ‘ (ππΊ, π4) (25)
For known solution temperature and pressure, the saturationconcentration can be deduced from solution property rela-tions. Following equations fix the properties of exiting vaporat (7)
π7 = ππΊ (26)
π7 = ππππΏβπ ππ‘ (π7, π3) (27)
β7 = βπ (π7, π7) (28)
(ii) Solution Heat Exchanger
Besides the trivial relations
π5 = π4π3 = π2
(29)
Mass and energy balance equations write
π5 = π4π3 = π2οΏ½οΏ½3 = οΏ½οΏ½2οΏ½οΏ½5 = οΏ½οΏ½4
(30)
β3 = β2 + οΏ½οΏ½4οΏ½οΏ½2
(β4 β β5) (31)
Considering the heat exchanger effectiveness, ππ»π, we havethe following further relations:
π5 = ππ»ππ2 + (1 β ππ»π) π4 (32)
β5 = βπππΏβπ ππ‘ (π5, π5) (33)
Journal of Engineering 5
β3 = βπππΏβπ ππ‘ (π3, π3) (34)
(iii) Solution Valve
Through the solution valve, the pressure is reduced fromcondenser to evaporator pressure. In addition to the usualmass balance-equations (π6 = π5) and (οΏ½οΏ½6 = οΏ½οΏ½5) we havethe relations
β6 = β5 (35)
π6 = ππππΏβπ ππ‘ (π4, β6) (36)
(iv) Solution Pump
Again, we have the trivial mass balances οΏ½οΏ½2 = οΏ½οΏ½1and π2 =π1. As for the water-pump, the pumping process is assumedisothermal (π2 = π1). During pumping, the enthalpy of therefrigerant-rich solution from absorber is increased by [(π2 βπ1)/π2], with [π2 = ππππΏβπ ππ‘(π2, β2)],
β2 = β1 + π2 β π1π2 (37)
(v) Absorber
Per definition, (ππ΄π΅ = π1) and (ππ΄π΅ = π1). For the liquidsolution (1) exiting the absorber we get in addition to themass and energy balance equations
οΏ½οΏ½1 = οΏ½οΏ½6 + οΏ½οΏ½11οΏ½οΏ½1π1 = οΏ½οΏ½6π6 (38)
οΏ½οΏ½π΄π΅ = (οΏ½οΏ½11β11 + οΏ½οΏ½6β6) β οΏ½οΏ½1β1 (39)
The property relations are
π1 = ππππΏβπ ππ (π1, π1) (40)
β1 = βπππΏβπ ππ‘ (π1, π1) (41)
3.3. Refrigerant Loop
(i) Condenser
Streams (8) and (15) flow in the condenser where theycondensate. Condensing temperature and pressure are ππΆπ· =π9 and ππΆπ· = π9, respectively. The mass and energy balancesaround the condenser write
οΏ½οΏ½9 = οΏ½οΏ½8 + οΏ½οΏ½15 = οΏ½οΏ½8 + οΏ½οΏ½19 = οΏ½οΏ½7 (42)
οΏ½οΏ½πΆπ· = οΏ½οΏ½9 (β8 β β9) (43)
Knowing the condensation temperature π9, pressure π9 aswell as the enthalpy of exiting liquid can be deduced as
π9 = ππβπ ππ‘ (π9) (44)
β9 = βπβπ ππ‘ (π9, π9 = 0) (45)
(ii) Refrigerant Expansion Valve
Liquid refrigerant (9) undergoes a pressure reduction beforeit enters the evaporator. Evaporation temperature and pres-sure are ππΈπ = π11 = π10 and ππΈπ = π10, respectively.
For fixed evaporator temperature ππΈπ and assumingsaturated vapor at exit, we can write ππΈπ = ππβπ ππ‘ (ππΈπ).
The mass and energy balances for the valve write
β10 = β9 (46)
οΏ½οΏ½10 = οΏ½οΏ½9 (47)
(iii) Evaporator
The evaporator equations are
οΏ½οΏ½πΈπ = οΏ½οΏ½11 (β11 β β10) (48)
οΏ½οΏ½11 = οΏ½οΏ½10
β11 = βπβπ ππ‘ (ππΈπ, π11 = 1) (49)
The πΆππβπ¦ππππ of the proposed absorption system, whenneglecting all pump work, can be expressed as
πΆππβπ¦ππππ=οΏ½οΏ½πΈποΏ½οΏ½ππΊ
(50)
4. Ejector 1D Model and Analysis
Because the performances of the proposed cycle dependlargely on ejector performances, a reliable ejector model isnecessary for the cycle simulations. In this paper, the ejectoris modelled basing on the 1D analyses in [25, 26].
In this type of model, it is assumed that
(i) primary fluid expands isentropically in nozzle, andthe exiting flow compresses isentropically in diffuser
(ii) inlet velocities of primary and entrained fluids areinsignificant
(iii) velocity of the compressed mixture at ejector outlet isneglected
(iv) mixing of primary and secondary fluids in the suctionchamber occurs at constant pressure
(v) flow in ejector is adiabatic
Isentropic efficiencies are introduced in the model to accountfor eventual irreversibility in the expansion process in pri-mary nozzle, (ππ), in the mixing process of primary andsecondary flow in themixing chamber, (ππ), and finally in thecompression process in the diffuser, (ππ). For the numericalsimulations we set ππ = 0.95, ππ = 0.95, and ππ = 1.
6 Journal of Engineering
4.1. Primary Nozzle. In the nozzle, the primary vapor (18)expands and accelerates. The Mach number π18π of the fluidat nozzle outlet plane (π), deduced fromenergy balance, writes
π18π = β 2πππΎ β 1 ((π18ππ
)(πΎβ1)/πΎ β 1) (51)
In this equation, ππ is the isentropic nozzle efficiency,defined as the ratio between actual enthalpy change andenthalpy change undergone during an isentropic process.
The expression for (π΄ π/π΄ π‘) the area ratio at nozzle throatand outlet is
π΄ ππ΄ π‘
= β 1π218π
( 2πΎ + 1 (1 + πΎ β 12 π218π))(πΎ+1)/(πΎβ1)
(52)
4.2. Suction Chamber. Because ππ < π19, the secondary fluid(19) expands in the suction chamber and is entrained bythe high-speed primary flow. The Mach number π19π of theentrained fluid at nozzle exit plane writes
π19π = β 2πΎ β 1 ((π19ππ
)(πΎβ1)/πΎ β 1) (53)
4.3.MixingChamber. Here, primary and secondary fluids aremixed.Theproperties of the resulting streamat section (π) arededuced from continuity, momentum, and energy equationsand expressed as function of the critical Mach numberπβ
π ,
πβπ = ππ πβ
18π + ππβ19πβπβ(1 + ππ) (1 + π) (54)
As can be noticed, themixtureπβπ is written as a combination
of critical Mach numbers of the original streams, πβ18π andπβ
19π. π in this equation stands for the temperature ratio ofincoming streams (19) and (18):
π = π19π18
(55)
The relationship between π and πβ at any point of theejector is given by the equation
π = β 2πβ2
(πΎ + 1) β (πΎ β 1)πβ2(56)
By the end of the mixing chamber, a shock wave occurs atsection (π). The flow changes from supersonic to subsonicconditions, producing simultaneously a sudden rise in thestatic pressure. The relation between the Mach numberupstream and downstream of the shock wave is given by
ππ = β 2/ (πΎ β 1) +π2π(2πΎ/ (πΎ β 1))π2
π β 1 (57)
The corresponding pressure increase writes
ππππ
= ππππ
β 1 + (1/2)π2π (πΎ β 1)1 + (1/2)π2π(πΎ β 1) (58)
4.4. Diffuser. Theexpression of the pressure lift in the diffuseris
π12ππ
= (1 + 12πππ2π (πΎ β 1))πΎ/(πΎβ1)
(59)
The ejector area ratio (π΄ π‘/π΄π), i.e., the ratio of nozzlethroat area and diffuser constant area section, writes
π΄ π‘π΄π
= π12π18
( πππ12
)1/πΎ
β β1 β ( πππ12
)(πΎβ1)/πΎβ 1(1 + ππ) (1 + π)β β(πΎ + 1) / (πΎ β 1)(2/ (πΎ + 1))1/(πΎβ1)
(60)
5. Results and Discussion
The EES machine model program is run to thermodynam-ically analyze the proposed hybrid single-effect absorptionrefrigeration system. The thermophysical properties of LiBr-H2O solution are estimated using the software property data-and model-bank.
The simulations are performed for the conditions given inTable 1. Evaporator temperature ππΊ is set to 4βC, condensertemperature ππΆπ· to 37βC, and absorber temperature ππ΄π΅ to(ππΆπ·β2). Condenser and absorber are both supposed water-cooled. The cooling medium is processed thereafter in acooling tower.
5.1. Program and Machine Model Validation. The simulationprogram is first validated by comparing our simulationresults for a conventional single-effect cycle with the resultspublished by Somers (2009) [27] for the same operatingconditions: evaporator temperature,1.3βC; condenser andabsorber temperatures at 40.2βC and 32.7βC, respectively;effectiveness of solution heat exchanger, 0.5; mass flow rateof solution leaving absorber, 1 kg/s. As can be noticed whencomparing the results in columns 2 and 3 of Table 2, both setsof data are in very good agreement. Therefore, we can nowproceed to the simulations of the proposed hybrid cycle withsome confidence.
The next step was to validate the adequacy of theconventional model by comparing the predicted, calculatedperformance with experimental data reported in [28] con-cerning a large capacity LiBr-chiller. Two different sets ofoperating conditions are considered. As can be observedwhen studying columns 4 to 7 in Table 2, the calculated datais for both tests very close to the reported data in [28]. Finally,the proposed ejector configuration model is validated using
Journal of Engineering 7
Table 1: Simulation input data.
Parameter Value Variation rangeSteam generator pressure, πππΊ, bar 15 10β15Generator temperature,ππΊ,
βC 80 65β90Evaporator temperature,ππΈπ,
βC 4 2β12Condensation temperature,ππΆπ·,
βC 37 28β37Absorber temperature,ππ΄π΅,
βC ππΆπ· β 2Table 2: Program and machine model validation.
Data 1 [27] Present work Data 2 [28] Present work Data 3[28] Present work
ππΊ,βC 90 101.6 83ππΈπ,βC 1.3 5 12.3ππΆπ·,βC 40.2 43 42ππ΄π΅,βC 32.7 38.3 39οΏ½οΏ½πΊ, kW 14.95 15.00 1150 1143 1100 1105οΏ½οΏ½πΈπ, kW 10.77 10.80 843 842.5 842.7 842.5πΆππ 0.73 0.72 0.73 0.74 0.76 0.76π4, % 62.6 62 65.5 65.8 57.2 58.5π3, % 57.4 56.3 56.5 57.4 53.1 53.4
0.8 0.9 1.0 1.1 1.20.8
0.9
1.0
1.1
1.2
COPhybrid (exp)
COP h
ybrid
(theo
)
Figure 3: Hybrid cycle model validation basing on experimentaldata of ref. [29].
the only available experimental data found in the literature[29]. As represented in Figure 3, a fair agreement betweencalculated and reported data is noticed. Discrepancy mayhave its source in inaccuracy of experimental and/or toosimple ejector model (ideal gas behavior).
5.2. Comparison of Hybrid and Conventional Cycle Per-formances. For purpose of illustration, the chiller cycle isrepresented in Figure 4 in the usual Oldham-diagram and inthe water (π β β)βdiagram in Figure 5.
We now proceed to the comparison of the performancesof the proposed cycle and the conventional basic cycle
(without ejector) for varying machine generator called alsodesorber-temperature (Figure 6), condenser temperature(Figure 7), and evaporator temperature (Figure 8).
As depicted in Figures 6β8, the coefficient of performanceof the hybrid cycle is in all cases larger than the πΆππ of theconventional cycle for the same operating conditions.
However, this performance enhancement is restrictedto a specific interval of machine-generator temperature, asFigure 6 clearly shows. Outside this temperature interval,both cycles are practically equivalent. Figure 6 shows also thatwith growing desorber temperatureππΊ theπΆππβcurve of thehybrid cycle first exceeds that of the basic cycle, reaches amaximum than decreases gradually, and resumes the curveof the conventional cycle πΆππ. It is also worth noticingthat the πΆππ of the hybrid cycle under optimal conditionsapproaches the πΆππ of double-effect conventional cycle.
Figures 7 and 8 depict the evolution of the πΆππ ofboth cycles with condenser and evaporator temperature,respectively, for (π18 = 15 bar; π18 β 200βC). Note that π18
is the steam generator temperature, not the chiller desorbertemperature, the abscissa in Figures 6β14. Both πΆππ areexpectably decreasing in the first case and increasing in thesecond. πΆππβπ¦ππππ is always larger than πΆππ of conventionalcycle because the constant maintained desorber-temperatureis set to 80βC, i.e., in the favourable interval 70βCβ90βC.In conclusion of this section we notice that an ejectorincorporated in the hybrid cycle (i) improves the cycleperformances and (ii) the maximal πΆππ is reached at lowermachine generator temperature.
5.3. Performances of the Hybrid Cycle. The effect observedpreviously in Figure 6 (enhancement of the cycle perfor-mance due to the incorporation of ejector in the driving
8 Journal of Engineering
50 10 15 20 25 30 40 45 50 55 60 65 70 75 80 85 90 100
105
110
11595 12035
50
10
5432
1
0.5
P [kPa]
Evaporatorpressure
Condenser pressure
Des
orbe
r tem
pera
ture
Con
dens
er te
mpe
ratu
re
Abso
rber
tem
pera
ture
11
9
1
4
6
Pure water,
=0
Aqueous LiBr solution,
=45
%
=50
%
=55
%
=60
%
=65%
T [βC]
Figure 4: Chiller cycle representation in the Oldham-diagram (πππΊ β 200βC; ππΊ = 85βC; ππΈπ = 4βC; ππΆπ· = 37βC).
0 200
104
103
102
101
100
400 600 800 1000 1200 1400 1600 1800 2000 2200 2400 2600 2800 3000
13
159
10 11
17 18
19
12
P (kPa)
h (kJ/kg)i
Figure 5: Chiller cycle representation in the water (π β β)βdiagram (πππΊ β 200βC; ππΊ = 85βC; ππΈπ = 4βC; ππΆπ· = 37βC).
70 75 80 85 90 95
basic cyclehybrid cycle
0.4
0.5
0.6
0.7
0.8
0.9
1.0
COP
οΌοΌ οΌ = 4βCοΌοΌοΌ = 37βC
Generator Temperature (βC)
Figure 6: πΆππ of hybrid and conventional cycle vs. machinegenerator temperature,ππΊ(π18 = 15 bar; π18 β 200βC).
compartment of the machine) depends on the primary flowpressure πππΊ = π18 used to activate the ejector. Increasingthis pressure expands this effect in magnitude and amplitudeas Figure 9 shows: the higher the steam-generator pressure
28 30 32 34 36 38 40 42
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1.0
COP
basic cyclehybrid cycle
Condenser Temperature (βC)
οΌοΌ οΌ = 4βCοΌοΌ = 80βC
Figure 7: πΆππ of hybrid and conventional cycle vs. condensertemperature,ππΆπ·.
(and consequently temperature), the larger the machine-generator temperature range where the cycle performanceis improved, and the higher the maximum πΆππ that could
Journal of Engineering 9
basic cyclehybrid cycle
οΌοΌοΌ = 37βCοΌοΌ = 80βC
2 4 6 8 10 12 140Evaporator Temperature (βC)
0.4
0.5
0.6
0.7
0.8
0.9
1.0CO
P
Figure 8: πΆππ of hybrid and conventional cycle vs. evaporatortemperature,ππΈπ.
οΌοΌοΌ = 37βCοΌοΌ οΌ = 4βC
0.2
0.4
0.6
0.8
1.0
COP h
ybrid
75 80 85 90 9570Generator Temperature (βC)
οΌοΌοΌ = 10barοΌοΌοΌ = 12barοΌοΌοΌ = 13bar
οΌοΌοΌ = 14barοΌοΌοΌ = 15bar
Figure 9: πΆππβπ¦ππππ vs. ππΊ for various steam-generator tempera-tures, πππΊ.
be reached inside this interval. On the opposite, when thesteamgenerator pressureπππΊ is decreased to 10 bar, practicallyno improvement more of the cycle performance is observedunder the prevailing conditions.
Figure 10 depicts the evolution of πΆππβπ¦ππππ with ππΊ
by varying the condenser temperature, ππΆπ·. It is observedthat the typical pink curve of Figure 6 is expectedly shiftedto lower machine-generator temperatures (with lower con-denser temperature, less high desorber temperature is neededto activate the cycle) with however concomitantly increasedmaximal πΆππ and enlarged favorable temperature interval,where the cycle performance is improved.
οΌοΌοΌ = 15barοΌοΌ οΌ = 4βC
οΌοΌοΌ = 32βCοΌοΌοΌ = 34βCοΌοΌοΌ = 36βC
0.2
0.4
0.6
0.8
1.0
1.2
COP h
ybrid
65 70 75 80 85 90 9560Generator Temperature (βC)
Figure 10: πΆππβπ¦ππππ vs. ππΊ for varying condenser temperature,ππΆπ·.
οΌοΌοΌ = 15barοΌοΌοΌ = 37βC
οΌοΌ οΌ = 4βCοΌοΌ οΌ = 6βC
οΌοΌ οΌ = 8βCοΌοΌ οΌ = 10βC
65 70 75 80 85 90 9560Generator Temperature (βC)
0.2
0.4
0.6
0.8
1.0
1.2
COP h
ybrid
Figure 11: πΆππβπ¦ππππ vs. ππΊ for varying evaporator temperature,ππΈπ.
Similar effects are observed in Figure 11 depicting the evo-lution of πΆππβπ¦ππππ with ππΊ by varying evaporator tempera-ture. Here, the typical COPβimproved portion of the curve isshifted to lowerππΊβvalues when the evaporator temperatureis increased, a thermodynamically more favourable situation.The πΆππ of the hybrid cycle rises from 0.85 to 1.12 forgenerator temperature decreasing from 78βC to 67βC whenthe evaporator temperature increases from 4βC to 12βC.
5.4. Ejector Performance. The ejector model presented inSection 4 will help us interpret the represented simulationresults in Figures 7β11 and assess the beneficial effectβand
10 Journal of Engineering
οΌοΌοΌ = 37βCοΌοΌ οΌ = 4βC
οΌοΌοΌ = 10barοΌοΌοΌ = 12barοΌοΌοΌ = 13bar
οΌοΌοΌ = 14barοΌοΌοΌ = 15bar
70 80 90 10060Generator Temperature (βC)
0.0
0.1
0.2
0.3
0.4
0.5En
trai
nmen
t rat
io
Figure 12: π vs. ππΊ for various primary pressure πππΊ.
limitsβof integration of an external ejector loop to a con-ventional absorption cycle. We first investigate the relationbetween the performance of the incorporated ejector, i.e.,its entrainment ratio π, and significant absorption machineparameters, namely, desorber temperature ππΊ, evaporatortemperature ππΈπ, and condenser temperature ππΆπ·. Figure 12depicts the evolution of π with ππΊ. For a given primarypressure πππΊ, the entrainment ratio decreases monotonouslywith ππΊ and finally vanishes for a maximal value of thedesorber temperature; i.e., secondary flow (19) is no moreentrained inside the ejector.The ejector is then off-design andits geometry should be changed. Same behaviour of π vs. ππΊ
is noticed if the steam pressure πππΊ is increased. However,in this case the curve is shifted upwards to larger values ofπ; i.e., more secondary vapour is sucked in the ejector for agiven temperature ππΊ, and the limit value of ππΊ where theentrainment ration vanishes is pushed farther away.
Similar behaviour is observed in Figure 13, when forfixed primary pressure the condenser temperature (sec-ondary pressure) is varied. If the condensation temperatureis reduced (or alternatively enlarged), the entrainment ratiois also decreased (or increased, respectively). However, thecurves π vs. ππΊ for the various condenser temperatures allconverge to the same point on the temperature-axis whereπ vanishes. This temperature depends solely on the primarysteam pressure.
Finally, Figure 14 shows that the evaporator temperaturehas practically no effect on the ejector performance by fixedπππΊ and ππΆπ·, as all π vs. ππΊ for the various tested ππΈπ aresuperimposed.
According to the ejector model presented in Section 3of the present paper, the entrainment ratio depends on sixindependent parameters: nozzle area ratio, primary flowand secondary flow properties, and backpressure, i.e., π =π(π΄ π/π΄ π‘, π18, π18, π19, π19, π12). The results presented in the
οΌοΌοΌ = 15bar
οΌοΌοΌ = 28βCοΌοΌοΌ = 30βCοΌοΌοΌ = 32βC
οΌοΌοΌ = 34βCοΌοΌοΌ = 36βC
οΌοΌ οΌ = 4βC
70 80 90 10060Generator Temperature (βC)
0.0
0.1
0.2
0.3
0.4
0.5
Entr
ainm
ent r
atio
Figure 13: π vs. ππΊ for various condenser temperature ππΆπ·.
οΌοΌοΌ = 15barοΌοΌοΌ = 37βC
οΌοΌ οΌ = 4βCοΌοΌ οΌ = 6βCοΌοΌ οΌ = 8βC
οΌοΌ οΌ = 10βCοΌοΌ οΌ = 12βC
70 80 90 10060Generator Temperature (βC)
0.0
0.1
0.2
0.3
0.4
0.5
Entr
ainm
ent r
atio
Figure 14: π vs. ππΊ for various evaporator temperatureππΈπ.
foregoing sections are obtained for simulations with thespecific conditions: (i) constant ejector nozzle ratio set to(π΄ π/π΄ π‘) = 17.3; (ii) saturated ejector-driving steam; i.e., π18
and π18 are then no more both independent; (iii) pressureof secondary flow π19 equals condenser pressure, an inde-pendent parameter; (iv) temperature π19 of flow π19 is notan independent variable. It depends on the processes takingplace in rest of the absorption chiller and in particular on thebackpressure,π12, which is considered here as an independentparameter.
Journal of Engineering 11
0.4
0.2
0.0
5
10
15P18 [bar]
1.0
0.5P12 [bar]
Figure 15: Entrainment ratio vs. primary pressure, π18, and back-pressure, π12, for fixed nozzle area ratio, (π΄ π/π΄ π‘) = 17.3, andsecondary pressure, π19 = 0.0628 bar.
In summary, the entrainment ratio depends then on justthree parameters
π = π (π18, π19, π12) (61)
Figure 15 illustrates this dependency for a fixed secondarypressure, π19 = 0.0628 bar, as it is the case for the datadepicted in Figures 6, 7, and 12. For a constant driving-steampressure π18, π increases with falling backpressure, becomesa maximum, and decreases thereafter abruptly to zero. Moregenerally, on increasing the ejector backpressure by fixedejector geometry, a gradual reduction in entrainment ratiois induced. The maximal value of π is the larger; i.e., thegreater the πΆππ-improvement, the higher the π18 . Further,when π18 becomes larger, the interval of backpressure π12
(and hence, the range of π12 as well as the range of desorbertemperature,π4) where a chiller performance enhancement isexpected, expands. The pressure difference (π18 β π12) drivesthe ejector, and the difference (π19 β ππ), where ππ is thepressure at nozzle exit, drives the entrainment process (Eq.(36)). With increasing primary pressure, ππ rises and comescloser to the secondary flow pressure π19 . The suction of thesecondary flow into the mixing chamber declines graduallyand eventually vanishes for ππ = π19. Consequently, at thislimit reached for π18 = 18 bar, π falls to zero. The verticalisobar-plane π18 = 18 bar sets a geometrical limit to the usednozzle design.
The π = 0 plane limits also the 3D surface of Figure 15.The calculations show that the Mach number ππ of themixed stream is there equal to π18π, the Mach number ofthe primary flow at nozzle exit; i.e., the mixed gas mass flowrate reduces to that of the primary flow and practically nosecondary gas is entrained. This constitutes a higher limit forthe design of the ejector area ratio (π΄ π‘/π΄π), which comes thenvery close to the nozzle area ratio, (π΄ π‘/π΄ π). The maximumvalue ofπ is found for minimal values of backpressure. At thelimit, the Mach number of mixed gas ππ is the lowest andequals that of the entrained secondary flowπ19π.
The πΆππ curves represented in Figures 6β9 depict itsevolution when the effects of both the ejector and the single-effect absorption chiller are combined. By increasing thebackpressure and, consequently, the desorber temperature,the πΆππ tends first to increase as it does for a conventionalcycle. The entrainment ratio however is decreasing. Theresulting outcome is then first an increase of πΆππ and thena decrease after passing a maximum where opposed effectscancel each other.
6. Conclusion
A hybrid single-effect cycle with water lithium-bromide asworking fluid and activated by a steam-ejector loop is pro-posed and theoretically investigated. Mathematical models ofthe hybrid cycle and the ejector are detailed. Results showthat entrainment ratio of the ejector depends on activating-steam pressure, on condenser temperature, and only slightlyon evaporator temperature. For a fixed steam pressure, theπΆππ of the hybrid cycle first surpasses that of the corre-sponding conventional cycle when the desorber temperatureis increased, passes by a maximum, and then resumes theperformance of the basic cycle. The maximum πΆππ of anejector-activated cycle is obtained at lower temperaturesthan that of a conventional system and can reach that of adouble-effect basic scheme. The span of machine generatortemperature where the πΆππ is enhanced depends on theprimary ejector pressure: it is larger for higher pressure. Theentrainment ratio of the ejector is found to increase withthe steam pressure and to decrease with rising backpressure.However, the performance of the ejector is confined to a spe-cific region of the parameter-surface. Outside this domain,the entrainment ratio vanishes and the ejector is off-design.
Nomenclature
π΄: Areaπ΄ππ΄π‘: Nozzle area ratio (π΄ π/π΄ π‘)π΄ππ΄π‘: Ejector area ratio (π΄π/π΄ π‘)πΆππ: Coefficient of performanceβ: Specific enthalpy (kJ/kg)οΏ½οΏ½: Mass flow rate (kg/s)π: Mach numberπβ: Critical Mach numberπ: Pressure (bar)οΏ½οΏ½: Heat transfer rate (kW)π : Universal gas constant (kJ/(kg K))π: Temperature (βC)οΏ½οΏ½: Work transfer rate (kW)π: Steam quality
Greek Symbols
πΎ: Ratio of steam specific heats (πΆπ/πΆV)πHX: Heat exchanger effectivenessπ: Nozzle, mixing, and diffuser efficiency
12 Journal of Engineering
π: LiBr concentration in solution (mass. %)π: Density (kg/m3)π: π19/π18π: Entrainment ratio (οΏ½οΏ½19/οΏ½οΏ½18).Subscripts
π΄π΅: Absorberππ: Backpressureπ: Constant section area (ejector)πΆπ·: Condenserπ: Diffuser (ejector)πΈπ: EvaporatorπΊ: Generatorπ: Nozzle exit plane (ejector)π: Plane in mixing chamber (ejector)π: Shockwave planeπ: Mixing chamber (ejector)π: Nozzle (ejector)π ππ‘: SaturationπππΏ: SolutionππΊ: Steam generatorπ: Water1β19: Referred state points.
Data Availability
The data used to support the findings of this study areavailable from the corresponding author upon request.
Conflicts of Interest
The authors declare that they have no conflicts of interest.
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