Performance Evaluation of Refrigeration Units in Natural Gas Liquid

Research ArticlePerformance Evaluation of Refrigeration Units inNatural Gas Liquid Extraction Plant

Awajiogak Anthony Ujile and Dirina Amesi

Department of ChemicalPetrochemical Engineering Faculty of Engineering Rivers State University of Science and TechnologyPMB 5080 Port Harcourt 50000 Nigeria

Correspondence should be addressed to Awajiogak Anthony Ujile ogakujileyahoocom

Received 29 December 2013 Accepted 23 February 2014 Published 25 March 2014

Academic Editor Ahmet Z Sahin

Copyright copy 2014 A A Ujile and D Amesi This is an open access article distributed under the Creative Commons AttributionLicense which permits unrestricted use distribution and reproduction in any medium provided the original work is properlycited

This paper has applied thermodynamics principles to evaluate the reliability of 390m3hr natural gas processing plant Thethermodynamics equations were utilized in the evaluation characterization and numerical simulation of key process parametersin natural gas liquid extraction plant The results obtained show the comparison of the coefficient of performance compressionratio isentropic work actual work electrical power requirements cooling water consumption in intercoolers compressor poweroutput compressor capacity and isentropic volumetric and mechanical efficiency of the two-stage refrigeration unit with a flashgas economizer and these were compared with the designed specifications The second law of thermodynamics was applied inanalyzing the refrigeration unit and the result shows that exergetic losses or lost work due to irreversibility falls within operatinglimit that is less than 10 Similarly the performance of expansion turbine (expander) parameters was monitored and the resultsindicate a considerable decrease in turbine efficiencies as the inlet gas pressure increases resulting in an increased power output ofthe turbine leading to a higher liquefaction rate

1 Introduction

The production and availability of natural gas liquid dependlargely on the supply of raw natural gas from productionwellhead and the operating conditions of the process unit thatmake up the extraction plant Most gas processing plants arefaced with problems ranging from inadequate supply poorfacility performance and human factorsThese problems canlead to low productivity of natural gas liquids and reductionin gas quality which could result in shutting down of theplant Poor facilities such as inadequate electricity supplyand processed water supply used in process equipment alsolead to intermittent operations and malfunction of processequipment such as pumps compressors and valves if they arenot adequately checked

Human factors may also result from the inability ofgas plant operators to monitor thermodynamics parameterssuch as the pressure flow rates and temperature on processequipment which could result in loss of data in control roomand unintended activation or deactivation of process devicesand reduce the plant efficiency In spite of the fact that these

thermodynamics parameters are monitored daily in a gasplant there are problems of low gas feed inlet pressure andinsufficient gas flow rateThese have resulted in low volume ofnatural gas liquids produced the extracted natural gas liquidsnot attaining the expected cryogenic temperature require-ment and variation in gas quality discharging from the outletof natural gas liquid extraction unit bottom product

If the available inlet gas pressure is low it can resultin compressor system suction pressure falling below atmo-spheric pressure This can also lead to air leakages into thecompressor system contributing to pulsation corrosion andlow heating value of natural gas In order to solve these prob-lems performance evaluation of the process units in naturalgas liquids extraction using thermodynamics principles isnecessary to ensure that these problems are minimized

It is seen from literature survey that several papers havebeen published focusing on thermodynamic analysis of a gasturbine power plant

Rahman et al [1] and Taniguchi andMiyamae [2] carriedout the study on the effects of ambient temperature andambient pressure as well as temperature of exhaust gases on

Hindawi Publishing CorporationJournal of ermodynamicsVolume 2014 Article ID 863408 7 pageshttpdxdoiorg1011552014863408

2 Journal of Thermodynamics

performance of gas turbine Khaliq and Kaushik [3] studiedthe efficiency of a gas turbine cogeneration system with heatrecovery steam generator

Keith and Kenneth [4] developed a new method ofapplying overall plantmaterial balance equation to determinethe performance of a natural gas processing plant Jibril et al[5] studied the simulation of expansion turbine (expander)for recovery of natural gas liquids from natural gas streamgas using HYPRO TECHrsquos HYSYS process simulation soft-ware Donnelly et al [6] carried out researches on processsimulation and optimization of cryogenic operations usingmultistream brazed aluminum heat exchangers Ganapathyet al [7] studied the energy analysis of operating lignite firedthermal power plant Design methodology for parametricstudy and thermodynamic performance evaluation of naturalgas process plant has been developed Refrigeration unit andexpansion turbine unit which are the major componentsof the plant were evaluated The expansion of natural gasthrough an expansion turbine with efficiencies shows a verygood performance Ujile and Alawa [8] The comparisonbetween the energy losses and the energy losses of thecomponents unit were investigated

This paper examined the reliability of a gas plant whichis the ability of a plant to maintain a stable efficiency withrespect to time using thermodynamics equations

2 MethodologySystem Descriptions

21 Propane Refrigeration Unit The simulation of processunits in propane refrigeration cycle involves the applicationsof thermodynamics principles to the following

(i) Calculate the amount of heat added to or removedfrom process streams

(ii) Estimate the power requirements for process equip-ment such as pumps compressors and turbine

(iii) Evaluate the performance of a flash separator atvarious temperatures and pressures

(iv) Determine the bubble and dew point temperatureassociated with distillation and bottom products

A systematic diagram of propane refrigeration cycle isshown in Figure 1 The refrigerant passes through the scrub-ber into the first stage compressor at 19 bar as saturated vaporwhere it mixes up with the vapor from the economizer unitIt is compressed to the inlet of the second stage compressorat 25 bar and discharges at 50 bar into the condenser assuperheated vapor It leaves the condenser at 47 bar andenters the first expansion valve and from the expansion valveinto an economizer drum where it flashes at an intermediatepressure of 33 bar The vapour phase with the economizer atthe top mixes up the vapour from the first stage compressorwhile the liquid discharges at the bottom and enters thesecond expansion valve into the chiller at a pressure of 25 barThe process gas (natural gas) flows inside the tubes connectedto the chiller at 41∘Cand countercurrently exchanges heat andgives up its energywith the liquid refrigerant surrounding thetubes The refrigerant boils up and leaves the chiller space assaturated vapor to be recompressed again

22 Expansion Turbine Unit A turbo expander or an expan-sion turbine is a centrifugal or axial flow turbine throughwhich a high pressure gas is expanded to produce useful workthat is used to drive a compressor Because work is extractedfrom the expanding high pressure gas the expansion is anapproximation by an isentropic process (a constant entropyprocess)

In the process of producing work the expander lowersthe bulk stream temperature which could result in partialliquefaction of the bulk stream (natural gas)

A systematic diagram of an expansion turbine (expander)is presented in Figure 2 The feed gas enters the gas treatingunit after which it is cooled firstly in a gas-to-gas heatexchanger and secondly the propane chiller The condensedliquid is removed in a separator and the gas from theseparator is cooled further in the low temperature gas-to-gasheat exchanger and fed into a second cooled separator Gasfrom the cooled separator expands through the expansionturbine to the pressure at the top of the demethanizer whichvaries from 100 to 450 bar During the expansion condensateis formed and the expander lowers the pressure of the inlet gasvalue to the demethanizer pressure of the range 100ndash450 bar

Thermodynamics Performance Parameters of Propane Refrig-eration UnitThe relevant parameters required for thermody-namics performance evaluation of propane refrigeration ofnatural gas may be considered as follows

Coefficient of Performance This is defined as the ratio of theheat absorbed (119876in) by the refrigerant while passing throughthe chiller to the work input (119882in) required compressingthe refrigerant in the compressor This is mathematicallyexpressed as [9 10]

COP = 119876in119882in (1)

but

119882in = 119882comp1 +119882comp2

= (1 minus 119883119866) (119867119861minus 119867119860) + 119872

119862(119867119864minus 119867119863MIX)

COP =(1 minus 119883

119866)119872119862(119867119860minus 119867119868)

(1 minus 119883119866) (119867119861minus 119867119860) + 119872

119862(119867119864minus 119867119863MIX)

(2)

Refrigeration Capacity The refrigeration capacity determinesthe rate of circulation of the refrigerant which in turndetermines the design and size of various units such ascondenser compressor evaporator (chiller) and expansionvalves This is expressed by the following equation

119877119888=119872119862(119867119860minus 119867119868)

210KJmin (3)

The mass flow rate of cooling water is calculated as follows

119872CW =119876

119862119901CWΔ119879 (4)

The overall performance of a compressor is affected mainlyby the inlet pressure and the interstage cooler efficiency

Journal of Thermodynamics 3

Economizer To econ

Econ liqChiller in

Chiller-Q

Chiller out

Comp1Mixer

Mixout

Comp2Comp2-hpComp1-hp

Condenser Condout

Comp1out Cond-Q

Chiller

19 bar 25 bar 50bar

33bar

27 barJT-2

JT-1

Figure 1 Process flow diagram of propane refrigeration of natural gas with application of Hysys software

(1) (2)

(3)First gasgas exchanger

Inletgas

Gastreating

Side exchanger

Second gasgas exchanger

Exp-compressor

Expander

JT valve

Recompressor

ReboilerD

emet

hani

zer

Demethanizedplant product

Residue gasto sales

Expa

nder

inle

t se

para

tor

Figure 2 Flow diagram of expander process

The simulation of process units in propane refrigerationresulted in loss of energy on the system The second avail-ability balance was applied to calculate the energy losses orlost work due to irreversibility of the process The thermody-namics equations applied on each of the process units are asfollows [10]

Compressor unit

119871119882= 119872119862[(119867119860minus 119879119900119878119860)] minus (119867

119861minus 119879119900119878119861) minus 119882

119864 (5)

condenser unit119871119882= 119879119900[119872119862(119878119862minus 119878119861)] + 119872CW119878CW(out) minus 119878CW(in)

119872CW = 119872119862 [119867119864minus 119867119865

119867CW(out) minus 119867CW(in)]

(6)

expansion valve

119871119882= 119872119862119879119900(119878119861minus 119878119862)

119871119882= 119872119862119879119900[(119878119860minus 119878119863) minus (119867119860minus 119867119863

119879119894

)] (7)

Refrigerant chiller the chiller is the unit where theprocess gas (natural gas) gives up its energy to the liquidrefrigerant The refrigerating effect which is the amount ofheat absorbed by refrigerant or removed from the refrigeratedspace is expressed as

119876in = (1 minus 119883119866)119872119862 (119867119860 minus 119867119868) (8)

The fraction of refrigerant vaporized in the chiller isdetermined from enthalpy balance as follows

119872119862(119867119867)119871= 119872119862119883119881(119867119868) + 119872

119862(1 minus 119883V) (119867119868)119871 (9)

where119883V is themole fraction of refrigerant that evaporates asit throttles to the economizer unit

Expansion Turbine Unit The relevant thermodynamicsparameters applied in expansion turbine unit in a gas plantare as follows

(i) Isentropic Turbine Efficiency This is defined as theratio of turbine actual work output to the work outputthatwould be achieved if the process between the inletand outlet pressure was isentropic By definition [11]

120578119864=ℎ1minus ℎ2119886

ℎ1minus ℎ2119878

(10)

(ii) Isentropic and Actual Work The isentropic work of aturbine is the work done by the turbine during theconstant entropy whereas actual work is the grosswork of a turbine

119882act = 119898 (ℎ2 minus ℎ1)

119882is = minus119898 (ℎ1 minus ℎ2119878) (11)


A

BD

EF

GH

I

m1 m2

Pres

sure

(bar

)

Enthalpy (kJkg)

Saturat

ed liq

uid

Expa

nsio

n

Condenser

Economizer

Chiller

Satu

rated

vapo

ur

1st sta

ge compres

sion

2nd st

age c

ompre

ssion

50bar47bar

33bar

19 bar57546

630

0

2704

2729

27167

27164

274551

valv

e 2

Expa

nsio

n va

lve 1

Figure 3 Pressure enthalpy diagram of a propane refrigeration cycle

(iii) Power Recovery The expander gas power recoveryis the output power of the turbine list to drive thecompressor By definition

119875119892= minus120578119898(ℎ1minus ℎ2119878) (12)

(iv) Theoretical and Actual Discharge Temperature Thedischarge and actual temperatures of a turbine areexpressed by the following equations [11 12]

1198791015840

2= 1198791(1199012

1199011

)

(119896minus1)119896

1198792= 1198791+ 1198791[(1199012

1199011

)

(119870minus1)119870

minus 1] 120578119864

(13)

The results obtained from the evaluation with the aboveequations are shown in graphical form in Figures 3 and 4

3 Results and Discussion

The results on the refrigeration unit of a gas plant wereobtained using the 119875-119878 and 119879-119878 diagrams

The two-stage compression refrigeration system in Fig-ures 3 and 4 operates with a pressure range of 19 and50 bar The refrigerant enters the first stage compressor at asuction pressure of 19 bar with enthalpy and entropy valuesof 2704KJkg and 7145 KJkgsdotK respectively It dischargesinto a mixer where it mixes up with the vapor leaving thetop of the flash economizer The refrigerant is compressedto the second stage compressor operating with a suctionpressure of 25 bar with enthalpy and entropy values of27164 KJkg and 7052KJkgsdotK respectively and discharges

at the condenser inlet pressure of 50 bar whereminus173867 Jkgof heat is rejected to the surrounding The enthalpy andentropy values of the condenser as a saturated vapour ata pressure of 47 bar were observed to be 630KJkg and18363KJkgsdotK respectively The refrigerant is throttled toa flash economizer at an intermediate pressure of 33 barDuring flashing 00034 kgs of vapor evaporates from theeconomizer drum while 01305 kgs of liquid is throttled tothe chiller at a pressure of 19 bar with enthalpy and entropyvalues of 5755 KJkg and 18545KJkgsdotK respectively Theprocess gas (natural gas) countercurrently exchanges heat andgives up its energy to the liquid refrigerant at a temperatureof 41∘CThe result of the two-stage compression refrigerationis summarized in Table 1

The process parameters of other auxiliary units inpropane refrigeration plant are shown in Table 2

The energy balance of the propane refrigeration unitobtained from the evaluation based on the equations high-lighted for the analysis is shown in Table 3

By the design conditions the centrifugal compressorsused in compressing the refrigerant should operate witha pressure ratio of 12 1 and 14 1 isentropic efficiency of70ndash80 volumetric efficiency of 60ndash89 and mechanicalefficiency of 20ndash50 By analytical method the pressure ratiowas found to be 1316 and 1414 the isentropic efficiency wasfound to be 4856 and 5197 the volumetric efficiency wasfound to be 550 and 467 for the first and second stagecompressors respectively while the mechanical efficienciesare 577 and 4651 respectivelyThe refrigeration cycle wasoperating with an overall coefficient of performance of 6237with a refrigerating capacity of 49222 tons after examiningthe performance of other auxiliary units within the systemsTables 1 2 and 3 show values obtained from the evaluation


A

B

C D

E

F

GH

I

Entropy (kJkg)

Satu

rate

d liq

uid

Expansionvalve 2

Expansionvalve 1

Condenser

Economizer

ChillerSaturated vapour

1st

com

pres

sion

stage

2nd

com

pres

sion

Stag

e

SA = 7145kJkg KhA = 2704kJkg

SB = 7052kJkg KhB = 27164kJkg

SE = 6822 kJkg KhE = 27485 kJkg

SI = 18545hI = 5755

SG = 18368hH = 5755

SF = 18368

1518

137

119

Tem

pera

ture

T(∘

C)

Figure 4 Temperature entropy diagram of a propane refrigeration cycle

Table 1 Summary of two-stage compression refrigeration

Parameters Units 1st stage compressor 2nd stage compressorSuction pressure bar 19 25Discharge pressure bar 25 50Compression ratio mdash 1316 1414Isentropic work of compression Jkg 2064425 2209491Actual work of compression Jkg 2064425 2187285Isentropic efficiency 4856 5198Mechanical efficiency 2071 4651Volumetric efficiency 550 467Actual power requirement KW 00969 00960Cooling water consumption in intercoolers m3s 72612 19282Compressor capacity m3s 21566 21096Work input in comp 1 and comp 2 Jkg 27314 27314Work output Jkg 119789 241411

The Kelvin-Planck statement of the second law of thermody-namics tells us that it is impossible to have a heat engine thatwill convert all the heat received from the high temperaturesource 119876

119867 into useful work in a thermodynamic cycle It

is necessary to reject part of the received heat to the lowtemperature source 119876

119871 In other words it is impossible to

have a 100 efficiency heat engine as corroborated by Kumaret al [13] and Jose and Simoes-Moreira [14]

4 Conclusion

The thermodynamics equations were applied to constructFigures 3 and 4 which determine the performance of theplant The recorded efficiencies ranging from 6392 to77 have shown that the overall performance has deviatedfrom the design efficiency This condition may increase theoperational cost of the plant

The following recommendations are highlighted toensure optimum efficiency and reliability of NGL plant

(i) The feed gas must be free from CO2and water This

affects plant efficiency and operations if not properlychecked by freezing the fittings valves and otherassociated equipment

(ii) The inlet strainer differential pressure must not behigh otherwise the expander will trip off on highdifferential pressure

(iii) The refrigeration unit must be operated within theoperating and design conditions to avoid overfreezingor warming of the demethanizer column

(iv) There is need for proper insulation of piping systemwithin the NGL extraction plant The reason is notto allow the surrounding heat to enter the system


Table 2 Process parameters of other auxilliary units

Parameters Unit Condenser Economizer Expansion valve 1 Expansion valve 2 Refrigerant chillerSuction pressure bar 50 33 47 33 23Discharge pressure bar 47 mdash 33 19 19Heat load Jkg minus173867 573378 mdash mdash 1703604Mass flow rate ofvapormdasheconomizer kgs mdash 00034 mdash mdash mdash

Mass flow rate ofliquidmdasheconomizer kgs mdash 01305 mdash mdash mdash

Mass flow rate of coolingwatermdashcondenser kgs 672 mdash mdash mdash mdash

Joule Thompson coefficient mdash mdash mdash 857 1286 mdashMole fraction of liquid remainingin chiller mdash 0035 000014

Mole fraction of vaporevaporated in chiller mdash 0025 0025 mdash 0045

Mass flow rate ofliquidmdashcondenser accumulator kgs 01339

Table 3 Energy balance propane refrigeration unit

Process units Energy gain (KJkg) Energy lost (KJkg) of lost workCompressor 1 1230 minus33428 minus0015Compressor 2 3211 minus50145 minus0023Condenser mdash 1248389 006Expansion valve JT-1 mdash 1249554 006Expansion valve JT-2 mdash minus14355478 662Economizer drum 573378 mdash mdashRefrigerant chiller 1703604 minus124271043 minus573Total 2281423 21698107 0972

thereby warming the system and there may be freez-ing out of the plant

(v) Proper sizing of the process line using the various linebalance methods to compare the amount of naturalgas entering the process plant with the amount put inshould be adopted at the design stage of the processplant

Nomenclature

119862119901CW The specific heat capacity ofwater

119867119860119867119861119867119863mix119867119864 and119867119868 The enthalpies of refrigerant

at compressor inlet (KJkg)outlet (KJkg) vapor mixturecoming from compressor andeconomizer units (KJkg) atthe condenser inlet (KJkg)and at the inlet of therefrigerant chiller (KJkg)respectively

ℎ1 ℎ2119886 ℎ2119878 The enthalpies of the gas

expander inlet gas expanderoutlet and at the exit pressurebut at the inlet entropy(KJkg) respectively

119896 The specific heat capacity ratioof gas to expander

119871119908 The lost work or rate of

irreversibility of refrigerant(KJkg)

119898 Themass flow rate of the gas toexpander (kgs)

119872119888 The refrigerant circulation rate119872CW The mass flow rate of cooling

water (kgs)1198751and 119875

2 The inlet and outlet gas

pressure to expander (bar)respectively

119876 The amount of heat removed inthe intercoolers

119878119860 119878119861 119878119862 119878119863 The entropies of refrigerant at

compressor inlet (KJkgsdotK)compressor outlet (KJkgsdotK)leaving the economizer(KJkgsdotK) and vapor mixture(KJkgsdotK) respectively

119878CW(in) 119878CW(out) The entropies of inlet coolingwater and outlet cooling waterrespectively

119879119900 The dead state temperature


1198791 The inlet gas temperature toexpander (∘C)

1198792 The actual gas temperature atexpander outlet (∘C)

1198791015840

2 The theoretical gas tempera-ture at expander outlet (∘C)

119882119864 The electrical power require-ment to compressor (KW)

119883119866 The mole fraction of vaporleaving the economizer unit

Δ119879 The temperature differencewithin the system boundary

120578119864 The expander efficiency120578119898 The mechanical efficiency

Conflict of Interests

The authors declare that there is no conflict of interestsregarding the publication of this paper

References

[1] M M Rahman T K Ibrahim K Kadirgama R Mamat andR A Bakar ldquoInfluence of operation conditions and ambienttemperature on performance of gas turbine power plantrdquoAdvanced Materials Research vol 189ndash193 pp 3007ndash3013 2011

[2] H Taniguchi and S Miyamae ldquoPower generation analysisfor high-temperature gas turbine in thermodynamic processrdquoJournal of Propulsion and Power vol 16 no 4 pp 557ndash561 2000

[3] A Khaliq and S C Kaushik ldquoThermodynamic performanceevaluation of combustion gas turbine cogeneration system withreheatrdquo Applied Thermal Engineering vol 24 no 13 pp 1785ndash1795 2004

[4] A B Keith and R H Kenneth Economic optimization ofnatural gas processing plant including business aspects [PhDdissertation] Texas A ampM University 2006

[5] K L Jibril A I Al-Humaizi A A Idriss and A A IbrahildquoSimulation of turbo-expander processes for recovering ofnatural gas liquids from natural gasrdquo Saudi Aramco Journal ofTechnology pp 9ndash14 2005

[6] S T Donnelly J C Polasek and J A Bullin Process Simulationand Optimization of Cryogenic Operations Using Multi-StreamBrazed Aluminium Exchangers Bryan Research and Engineer-ing 2006

[7] T Ganapathy N Alagumurthi R P Gakkhar and K Muruge-san ldquoExergy analysis of operating lignite fired thermal powerplantrdquo Journal of Engineering Science and Technology Reviewvol 2 no 1 pp 123ndash130 2009

[8] A A Ujile and L Alawa ldquoThermodynamic evaluation of a nat-ural gas processing plant case studymdashObiafuObrikom RiversState Nigeriardquo Journal of Advanced Science and EngineeringTechnology vol 2 no 1 pp 76ndash80 2012

[9] T D Eastop and A McConkey Applied Thermodynamics forEngineering Technologists Pearson Education 5th edition 2005

[10] Y A Cengel and M A BolesThermodynamics an EngineeringApproachMcGraw-Hill NewYork NYUSA 4th edition 2002

[11] I S Stanley Chemical Biochemical and EngineeringThermody-namics John Wiley amp Sons 4th edition 2006

[12] C Arora Refrigeration and Air Conditioning McGraw HillDelhi India 5th edition 2005

[13] A Kumar S S Kachhwaha and R SMishra ldquoThermodynamicanalysis of a regenerative gas turbine cogeneration plantrdquoJournal of Scientific and Industrial Research vol 69 no 3 pp225ndash231 2010

[14] R Jose and Simoes-Moreira ldquoThermal Power Plant Perfor-mance Analysisrdquo 2012 httpwwwspringercom978-1-4471-2308-8

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performance of gas turbine Khaliq and Kaushik [3] studiedthe efficiency of a gas turbine cogeneration system with heatrecovery steam generator

Keith and Kenneth [4] developed a new method ofapplying overall plantmaterial balance equation to determinethe performance of a natural gas processing plant Jibril et al[5] studied the simulation of expansion turbine (expander)for recovery of natural gas liquids from natural gas streamgas using HYPRO TECHrsquos HYSYS process simulation soft-ware Donnelly et al [6] carried out researches on processsimulation and optimization of cryogenic operations usingmultistream brazed aluminum heat exchangers Ganapathyet al [7] studied the energy analysis of operating lignite firedthermal power plant Design methodology for parametricstudy and thermodynamic performance evaluation of naturalgas process plant has been developed Refrigeration unit andexpansion turbine unit which are the major componentsof the plant were evaluated The expansion of natural gasthrough an expansion turbine with efficiencies shows a verygood performance Ujile and Alawa [8] The comparisonbetween the energy losses and the energy losses of thecomponents unit were investigated

This paper examined the reliability of a gas plant whichis the ability of a plant to maintain a stable efficiency withrespect to time using thermodynamics equations

2 MethodologySystem Descriptions

21 Propane Refrigeration Unit The simulation of processunits in propane refrigeration cycle involves the applicationsof thermodynamics principles to the following

(i) Calculate the amount of heat added to or removedfrom process streams

(ii) Estimate the power requirements for process equip-ment such as pumps compressors and turbine

(iii) Evaluate the performance of a flash separator atvarious temperatures and pressures

(iv) Determine the bubble and dew point temperatureassociated with distillation and bottom products

A systematic diagram of propane refrigeration cycle isshown in Figure 1 The refrigerant passes through the scrub-ber into the first stage compressor at 19 bar as saturated vaporwhere it mixes up with the vapor from the economizer unitIt is compressed to the inlet of the second stage compressorat 25 bar and discharges at 50 bar into the condenser assuperheated vapor It leaves the condenser at 47 bar andenters the first expansion valve and from the expansion valveinto an economizer drum where it flashes at an intermediatepressure of 33 bar The vapour phase with the economizer atthe top mixes up the vapour from the first stage compressorwhile the liquid discharges at the bottom and enters thesecond expansion valve into the chiller at a pressure of 25 barThe process gas (natural gas) flows inside the tubes connectedto the chiller at 41∘Cand countercurrently exchanges heat andgives up its energywith the liquid refrigerant surrounding thetubes The refrigerant boils up and leaves the chiller space assaturated vapor to be recompressed again

22 Expansion Turbine Unit A turbo expander or an expan-sion turbine is a centrifugal or axial flow turbine throughwhich a high pressure gas is expanded to produce useful workthat is used to drive a compressor Because work is extractedfrom the expanding high pressure gas the expansion is anapproximation by an isentropic process (a constant entropyprocess)

In the process of producing work the expander lowersthe bulk stream temperature which could result in partialliquefaction of the bulk stream (natural gas)

A systematic diagram of an expansion turbine (expander)is presented in Figure 2 The feed gas enters the gas treatingunit after which it is cooled firstly in a gas-to-gas heatexchanger and secondly the propane chiller The condensedliquid is removed in a separator and the gas from theseparator is cooled further in the low temperature gas-to-gasheat exchanger and fed into a second cooled separator Gasfrom the cooled separator expands through the expansionturbine to the pressure at the top of the demethanizer whichvaries from 100 to 450 bar During the expansion condensateis formed and the expander lowers the pressure of the inlet gasvalue to the demethanizer pressure of the range 100ndash450 bar

Thermodynamics Performance Parameters of Propane Refrig-eration UnitThe relevant parameters required for thermody-namics performance evaluation of propane refrigeration ofnatural gas may be considered as follows

Coefficient of Performance This is defined as the ratio of theheat absorbed (119876in) by the refrigerant while passing throughthe chiller to the work input (119882in) required compressingthe refrigerant in the compressor This is mathematicallyexpressed as [9 10]

COP = 119876in119882in (1)

but

119882in = 119882comp1 +119882comp2

= (1 minus 119883119866) (119867119861minus 119867119860) + 119872

119862(119867119864minus 119867119863MIX)

COP =(1 minus 119883

119866)119872119862(119867119860minus 119867119868)

(1 minus 119883119866) (119867119861minus 119867119860) + 119872

119862(119867119864minus 119867119863MIX)

(2)

Refrigeration Capacity The refrigeration capacity determinesthe rate of circulation of the refrigerant which in turndetermines the design and size of various units such ascondenser compressor evaporator (chiller) and expansionvalves This is expressed by the following equation

119877119888=119872119862(119867119860minus 119867119868)

210KJmin (3)

The mass flow rate of cooling water is calculated as follows

119872CW =119876

119862119901CWΔ119879 (4)

The overall performance of a compressor is affected mainlyby the inlet pressure and the interstage cooler efficiency


Economizer To econ

Econ liqChiller in

Chiller-Q

Chiller out

Comp1Mixer

Mixout


Condenser Condout

Comp1out Cond-Q

Chiller

19 bar 25 bar 50bar

33bar

27 barJT-2

JT-1


(1) (2)


Inletgas

Gastreating

Side exchanger


Exp-compressor

Expander

JT valve

Recompressor

ReboilerD

emet

hani

zer


Residue gasto sales

Expa

nder

inle

t se

para

tor



Compressor unit

119871119882= 119872119862[(119867119860minus 119879119900119878119860)] minus (119867

119861minus 119879119900119878119861) minus 119882

119864 (5)


119872CW = 119872119862 [119867119864minus 119867119865


(6)

expansion valve

119871119882= 119872119862119879119900(119878119861minus 119878119862)


119879119894

)] (7)


119876in = (1 minus 119883119866)119872119862 (119867119860 minus 119867119868) (8)


119872119862(119867119867)119871= 119872119862119883119881(119867119868) + 119872

119862(1 minus 119883V) (119867119868)119871 (9)




120578119864=ℎ1minus ℎ2119886


(10)


119882act = 119898 (ℎ2 minus ℎ1)

119882is = minus119898 (ℎ1 minus ℎ2119878) (11)


A

BD

EF

GH

I

m1 m2

Pres

sure

(bar

)

Enthalpy (kJkg)

Saturat

ed liq

uid

Expa

nsio

n

Condenser

Economizer

Chiller

Satu

rated

vapo

ur

1st sta

ge compres

sion

2nd st

age c

ompre

ssion

50bar47bar

33bar

19 bar57546

630

0

2704

2729

27167

27164

274551

valv

e 2

Expa

nsio

n va

lve 1



119875119892= minus120578119898(ℎ1minus ℎ2119878) (12)


1198791015840

2= 1198791(1199012

1199011

)

(119896minus1)119896

1198792= 1198791+ 1198791[(1199012

1199011

)

(119870minus1)119870

minus 1] 120578119864

(13)










A

B

C D

E

F

GH

I

Entropy (kJkg)

Satu

rate

d liq

uid

Expansionvalve 2

Expansionvalve 1

Condenser

Economizer


1st

com

pres

sion

stage

2nd

com

pres

sion

Stag

e




SI = 18545hI = 5755

SG = 18368hH = 5755

SF = 18368

1518

137

119

Tem

pera

ture

T(∘

C)









4 Conclusion




















Nomenclature











water (kgs)1198751and 119875











1198791015840








References




















FluidsJournal of


Journal of










Superconductivity




GravityJournal of








Journal of






Volume 2014


PhotonicsJournal of


Journal of

Biophysics




Economizer To econ

Econ liqChiller in

Chiller-Q

Chiller out

Comp1Mixer

Mixout


Condenser Condout

Comp1out Cond-Q

Chiller

19 bar 25 bar 50bar

33bar

27 barJT-2

JT-1


(1) (2)


Inletgas

Gastreating

Side exchanger


Exp-compressor

Expander

JT valve

Recompressor

ReboilerD

emet

hani

zer


Residue gasto sales

Expa

nder

inle

t se

para

tor



Compressor unit

119871119882= 119872119862[(119867119860minus 119879119900119878119860)] minus (119867

119861minus 119879119900119878119861) minus 119882

119864 (5)


119872CW = 119872119862 [119867119864minus 119867119865


(6)

expansion valve

119871119882= 119872119862119879119900(119878119861minus 119878119862)


119879119894

)] (7)


119876in = (1 minus 119883119866)119872119862 (119867119860 minus 119867119868) (8)


119872119862(119867119867)119871= 119872119862119883119881(119867119868) + 119872

119862(1 minus 119883V) (119867119868)119871 (9)




120578119864=ℎ1minus ℎ2119886


(10)


119882act = 119898 (ℎ2 minus ℎ1)

119882is = minus119898 (ℎ1 minus ℎ2119878) (11)


A

BD

EF

GH

I

m1 m2

Pres

sure

(bar

)

Enthalpy (kJkg)

Saturat

ed liq

uid

Expa

nsio

n

Condenser

Economizer

Chiller

Satu

rated

vapo

ur

1st sta

ge compres

sion

2nd st

age c

ompre

ssion

50bar47bar

33bar

19 bar57546

630

0

2704

2729

27167

27164

274551

valv

e 2

Expa

nsio

n va

lve 1



119875119892= minus120578119898(ℎ1minus ℎ2119878) (12)


1198791015840

2= 1198791(1199012

1199011

)

(119896minus1)119896

1198792= 1198791+ 1198791[(1199012

1199011

)

(119870minus1)119870

minus 1] 120578119864

(13)










A

B

C D

E

F

GH

I

Entropy (kJkg)

Satu

rate

d liq

uid

Expansionvalve 2

Expansionvalve 1

Condenser

Economizer


1st

com

pres

sion

stage

2nd

com

pres

sion

Stag

e




SI = 18545hI = 5755

SG = 18368hH = 5755

SF = 18368

1518

137

119

Tem

pera

ture

T(∘

C)









4 Conclusion




















Nomenclature











water (kgs)1198751and 119875











1198791015840








References




















FluidsJournal of


Journal of










Superconductivity




GravityJournal of








Journal of






Volume 2014


PhotonicsJournal of


Journal of

Biophysics




A

BD

EF

GH

I

m1 m2

Pres

sure

(bar

)

Enthalpy (kJkg)

Saturat

ed liq

uid

Expa

nsio

n

Condenser

Economizer

Chiller

Satu

rated

vapo

ur

1st sta

ge compres

sion

2nd st

age c

ompre

ssion

50bar47bar

33bar

19 bar57546

630

0

2704

2729

27167

27164

274551

valv

e 2

Expa

nsio

n va

lve 1



119875119892= minus120578119898(ℎ1minus ℎ2119878) (12)


1198791015840

2= 1198791(1199012

1199011

)

(119896minus1)119896

1198792= 1198791+ 1198791[(1199012

1199011

)

(119870minus1)119870

minus 1] 120578119864

(13)










A

B

C D

E

F

GH

I

Entropy (kJkg)

Satu

rate

d liq

uid

Expansionvalve 2

Expansionvalve 1

Condenser

Economizer


1st

com

pres

sion

stage

2nd

com

pres

sion

Stag

e




SI = 18545hI = 5755

SG = 18368hH = 5755

SF = 18368

1518

137

119

Tem

pera

ture

T(∘

C)









4 Conclusion




















Nomenclature











water (kgs)1198751and 119875











1198791015840








References




















FluidsJournal of


Journal of










Superconductivity




GravityJournal of








Journal of






Volume 2014


PhotonicsJournal of


Journal of

Biophysics




A

B

C D

E

F

GH

I

Entropy (kJkg)

Satu

rate

d liq

uid

Expansionvalve 2

Expansionvalve 1

Condenser

Economizer


1st

com

pres

sion

stage

2nd

com

pres

sion

Stag

e




SI = 18545hI = 5755

SG = 18368hH = 5755

SF = 18368

1518

137

119

Tem

pera

ture

T(∘

C)









4 Conclusion




















Nomenclature











water (kgs)1198751and 119875











1198791015840








References




















FluidsJournal of


Journal of










Superconductivity




GravityJournal of








Journal of






Volume 2014


PhotonicsJournal of


Journal of

Biophysics















Nomenclature











water (kgs)1198751and 119875











1198791015840








References




















FluidsJournal of


Journal of










Superconductivity




GravityJournal of








Journal of






Volume 2014


PhotonicsJournal of


Journal of

Biophysics






1198791015840








References




















FluidsJournal of


Journal of










Superconductivity




GravityJournal of








Journal of






Volume 2014


PhotonicsJournal of


Journal of

Biophysics








FluidsJournal of


Journal of










Superconductivity




GravityJournal of








Journal of






Volume 2014


PhotonicsJournal of


Journal of

Biophysics



Documents

Performance Evaluation of Refrigeration Units in Natural Gas Liquid