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Applied Thermal Engineering 31 (2011) 4083e4090

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Applied Thermal Engineering

journal homepage: www.elsevier .com/locate/apthermeng

The search for optimum condenser cooling water flow rate in a thermalpower plant

A.N. Anozie, O.J. Odejobi*

Applied Thermodynamics and Process Design Research Laboratory, Department of Chemical Engineering, Obafemi Awolowo University, Ile-Ife, Nigeria

a r t i c l e i n f o

Article history:Received 14 January 2011Accepted 10 August 2011Available online 17 August 2011

Keywords:Cooling water flowrateHeat transfer areaThermal power plantHeat recoveryCycle efficiencyHeat exchanger network

* Corresponding author. Tel.: þ234 8060790337.E-mail address: oludareodejobi@yahoo.co.uk (O.J.

1359-4311/$ e see front matter � 2011 Elsevier Ltd.doi:10.1016/j.applthermaleng.2011.08.014

a b s t r a c t

Heat losses from the thermal power plant cycle are due mainly to heat rejection through the condenser.Operating the condenser at optimum circulation water flowrate is essentially important to ensuremaximum efficiency and minimum operating cost of the plant. In this study, computer program codeswere developed in Microsoft Excel macros for simulation of a thermal plant at various circulation waterflowrate, to determine the optimum condenser cooling water flowrate for the process. The studyrevealed that operating the condenser at reduced cooling water flow rate of 32,000 m3/h instead of thebase case scenario of 32,660 m3/h, reduced the total heat transfer area requirement from 13,256 m2 to8,113 m2, with the condenser making the highest contribution to heat transfer area reduction. Theannualized capital cost also reduced to $12,271,064.30/yr from $16,809,876.50/yr. There was 2% increasein the cycle efficiency and fuel saving of 3.8% was achieved. The economic implications of heat recoveryimprovement were modifications to the air ejector, gland condenser, and replacement of the drain cooler,low pressure heater and high pressure heaters. The fixed capital for plant modification was $4,694,220.96with payback period of 1.8 years.

� 2011 Elsevier Ltd. All rights reserved.

1. Introduction

Several studies have been made to optimize the design and/oroperation of thermal power plants. The underlying concepts,objective indices andmethods of optimization in these studies vary.The major concepts which have been applied in these studies arepinch technology, energy analysis, exergy analysis and exer-goeconomic analysis. The objective indices are usually economicand efficiency parameters; and the methods of optimization varyfrom advanced to simple methods.

Pinch technology represents a set of thermodynamically basedmethods that guarantee minimum energy levels in design of heatexchanger networks. Pinch technology and cycle efficiency target-ing have been used to increase the cycle efficiency of a steam powerplant and to reduce the fuel consumption of the plant [1]. Pinchanalysis was used to reduce energy penalty in a coal-fired powerplant with carbon capture; and cooling water requirement wasreduced compared to base case [2]. The pinch concept has also beenused as an aid in selecting the optimum combined heat and power(CHP) systems so that the overall energy consumption of the

Odejobi).

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process is minimized [3]; and to integrate a new process into anexisting site utility system for optimum performance [4].

Energy analysis is based on the first law of thermodynamics.The first law of thermodynamics is a conservation lawwhich statesthat energy cannot be destroyed but can be transformed from oneform to another. Energy analysis quantifies the flow of energy inprocesses. The first law efficiency of thermal power plants wasshown to increase slightly with work output, giving highest effi-ciencies at full load [5,6]. Contour plotting of first law thermalefficiency and reheat pressures for optimization of reheat regen-erative thermal power plants has been reported [7]. It has beenshown that varying the condenser cooling water flowrate haseffect on the cycle efficiency of a thermal power plant [8]. Energyanalysis and optimization of a thermal power plant using linearprogramming (LP) and EXCEL solver have been used to improvethe operation of a thermal plant [9]. Mixed integer linearprogramming (MILP) optimizationmethod has been used to derivethe cost region diagrams for energy systems with a condensingturbine [10].

Exergy analysis also called second-law analysis identifies themagnitudes and the locations of exergy losses, in order to improvethe existing systems, processes or components, or to develop newprocesses or systems [11,12]. Exergy analysis and parametric studyhave been used to make optimum design decisions in a logical

A.N. Anozie, O.J. Odejobi / Applied Thermal Engineering 31 (2011) 4083e40904084

manner in a thermal power plant [6]. The concept has been used tooptimize the first and second reheat pressures in a thermal powerplant [7]. A comparison has been made of coal-fired and nuclearsteam power plants using exergy analysis to identify areas withpotential for performance improvement [13]. A reduction inproduction and fuel consumption costs was achieved by exergyanalysis of a thermal power plant [14]. The exergy destruction ina combined Carnot cycle has been quantified [15]. Exergy analysisfor the optimization of cogeneration steam plants has been per-formed [16,17].

Exergoeconomic analysis is the second-law based economicanalysis and it uses exergy to apportion the production cost todifferent parts of the production routes. Exogoeconomic analysisanswers the questions of the cost of thermodynamics ineffi-ciency in processes and measures of improving the cost effec-tiveness of the overall process [18]. Exergoeconomics has beenapplied in the design of energy efficient systems [19,20]; in theanalysis of solar thermal power plant [21]; and in the optimi-zation of a combined cycle power plant [22]. Exergoeconomicanalysis and various optimization techniques have been appliedin the analysis of thermal systems. The classical optimizationmethod of Lagrange multipliers has been used by some workers[23,24]; and the multivariable optimization technique has beenapplied [25].

This work belongs to the class of energy analysis and optimi-zation of a thermal power plant. Specifically, it focused on thesearch for optimum condenser cooling water flowrate in a thermalpower plant using total cost (comprising cost of retrofitting theheat exchangers and energy cost) and cycle efficiency as objectiveindices, and using the simple method of graphs and tables todetermine the optimum value. Computer program codes based onheat and mass balances were developed using Microsoft Excelmacros for simulation of the thermal power plant and heat

Compressedair

5b

HPH6

49

HPH5

48

L

6

91

14

16

42

43

112

2

3

4

5a

7

8

10

11

15

44

45

46

4750

51

Fuel

Flue gas to stack

Furnace

BFP

Deaerator

Boiler

Reheater

HPT IPT

High Pressure Turbine (HPT); Intermediate Pressure Turbine (IPT); Low Pressure Turbin(LPH); High Pressure Heater (HPH); Condensate Extraction Pump (CEP); Boiler Feed W

Fig. 1. Egbin Thermal Power Pl

exchangers areas estimation and cost analysis at varying condensercooling water flowrate in order to determine the optimumcondenser cooling water flowrate for the plant. This work defersfrom our earlier work [8] in that the earlier work was not posed asan optimization problem but showed that cycle efficiency could beimproved by varying condenser cooling water flowrate.

The condenser in a thermal power plant generally uses eithercirculating cooling water from a cooling tower to reject waste heatto the atmosphere, or once-through water from a river, lake orocean. Since the greatest amount of heat lost in the thermal powerplant is due to heat rejection from the condenser [26], the energyand the fuel savings and subsequent improvement in the plantefficiency can be achieved by reducing heat rejection through thecondenser by reducing the circulating water flow rate. However,insufficient condenser cooling water may lead to a power limitationdue to vacuum losses in the condenser [27]. Hence, there is need todetermine optimum flowrate of the condenser cooling water forbest performance of the thermal plant and improved efficiency ofthe thermal cycle.

2. Process description

The case study is a thermal power plant, located at Egbin inIkorodu, Lagos State, Nigeria. The plant consists of six sets or unitseach of 220 MW. The sets are dual firing using either natural gasand/or high pour fuel oil (HPFO). Fig. 1 shows the flow diagram ofa complete set and the heat exchanger network. Each set has threeturbines, namely: high pressure turbine (HPT), intermediate pres-sure turbine (IPT) and low pressure turbine (LPT). The turbines aremounted on a single shaft and a generator is coupled directly withthem. Single stage reheating is employed between HPT and IPT. TheLPTexhaust gets condensed in the condenser. The condensed steam(stream 30) is pumped from the hot well by the condensate

Condenser

41

PH3

37

LPH1LPH38

17

18

2427

2622

23

21

20

193

29

28

30

31

32

33

34

35

39

36

40

25

Overall system

C.W out

C.W in

Generator

Drain Cooler

CE

Gland Condenser

Air ejector

G

SS

T3

Hot well

LPT

e (LPT) ; Steam Seal Regulator (SSR); Cooling Water (CW); Low Pressure Heater ater Pump (BFP); Stream Numbers (1,...,51).

ant Process flow diagram.

A.N. Anozie, O.J. Odejobi / Applied Thermal Engineering 31 (2011) 4083e4090 4085

extraction pump (CEP) into the steam air ejector. The water isfurther pumped through the gland condenser, the drain cooler andthe low pressure feed water heaters LPH1, LPH2 and LPH3. Thefeed water from the deaerator (stream 44) is pumped into thefurnace and high pressure heaters, HPH5 and HPH6 by theboiler feed pump (BFP). The extraction steam (streams 15 and 10)for the high pressure heaters are coming from the intermediatepressure turbine and the exhaust from high pressure turbinecascaded backward into the reheater. Superheated steam (stream1) from the boiler enters directly into high pressure turbine whereit provides shaft work that drives the turbine blade. Exhauststeam from high pressure turbine (stream 5b) goes back into thereheater before it passes on to intermediate pressure turbine(stream 12). The exhaust from the intermediate pressure turbine(stream 18) enters directly into the low pressure turbine to provideshaft work that drives the generator to generate electricity. Theexhaust steam (stream 27) from low pressure turbine is condensedin the condenser using Lagoon water as circulation water coolingsystem (CW).

Fig. 2. Algorithmfor theevaluationof unknownthermodynamicproperties andestimationof co

3. Methodology

3.1. Plant simulation and estimation of heat exchangers areas

In this work, process operating data were used. The plantcomponents were grouped into work modules. Inlet and outletstreams from the components in each module were extracted andentered into the Microsoft Excel worksheets. Computer programcodes were developed and written in Microsft Excel Macros for thesimulation of the entire plant for varying condenser cooling waterflowrate. The programme codes evaluate unknown thermodynamicproperties of the process streams through interaction betweenwork modules. Saturated water steam Tables for evaluation ofthermodynamic properties of the process streams in the course ofsimulationwere extracted from literature and incorporated into theprogramme codes to ease estimation of unknown thermodynamicproperties for the process streams at varying condenser coolingwater flowrate. A sample algorithm for the evaluation of unknownthermodynamic properties of the process streams for estimation of

ndenserandotherheatexchangerproperties for varyingcondenser coolingwaterflowrate.

Table 1Comparison of measured and calculated thermodynamic parameters.

Parameters Measuredvalue

Calculatedvalue

Condenser Temperature (�C) 42.1(�0.5) 42.1Air Ejector tube side inlet temperature (�C) 42.1(�0.5) 42.1Air Ejector tube side outlet temperature (�C) 42.2(�0.5) 42.7Gland condenser tube side inlet temperature (�C) 42.2 (�0.5) 42.7Gland condenser tube side outlet temperature (�C) 43.7 (�0.5) 43.9Drain cooler tube side inlet temperature (�C) 43.7 (�0.5) 43.9Drain cooler tube side outlet temperature (�C) 48.6 (�0.5) 49.6LPH 1 tube side inlet temperature (�C) 48.6 (�0.5) 49.6LPH 1 tube side outlet temperature (�C) 86.7 (�0.5) 86.7LPH 2 tube side inlet temperature (�C) 86.7 (�0.5) 86.7LPH 2 tube side outlet temperature (�C) 110 (�0.5) 110LPH 3 tube side inlet temperature (�C) 110 (�0.5) 110LPH 3 tube side outlet temperature (�C) 134.8 (�0.5) 134.2HPH 5 tube side inlet temperature (�C) 165.5 (�0.5) 165.5HPH 5 tube side outlet temperature (�C) 196.6 (�0.5) 196.6HPH 6 tube side inlet temperature (�C) 196.6 (�0.5) 196.6HPH 6 tube side outlet temperature (�C) 236.6 (�0.5) 236.6Condenser Pressure (kPa) 8.24 (�0.05) 8.30Air Ejector Pressure (kPa) 8.57 (�0.05) 8.59Gland Condenser Pressure (kPa) 9.09 (�0.05) 9.10Drain Cooler Pressure (kPa) 12.11 (�0.05) 12.13LPH 1 Pressure (kPa) 220.50 (�1.0) 231.70LPH 2 Pressure (kPa) 332.5 (�1.0) 343.50LPH 3 Pressure (kPa) 593.20 (�2.0) 599.90HPH 5 Pressure (kPa) 1754.00 (�2.0) 1761.00HPH 6 Pressure (kPa) 3085.00 (�2.0) 3087.00

A.N. Anozie, O.J. Odejobi / Applied Thermal Engineering 31 (2011) 4083e40904086

condenser and other heat exchanger properties for varyingcondenser cooling water flowrate is presented in Fig. 2.

The commonly used heat exchanger in this kind of system is theU-tube shell and tube. The duty of each heat exchanger in thenetwork was estimated by heat balance using Eqs. (1) and (2) [28].

Q ¼ WsðHsi � HsoÞ ¼ WT ðHTo � HTiÞ (1)

Q ¼ CPsðTsi � TsoÞ ¼ CPTðTTo � TTiÞ (2)

For the mixing of streams inlets to the shell side of the heatexchanger, the effective enthalpy is as given in Eq. (3) [26]:

Effective Enthalpy ¼Pn

0HjWjPn0Wj

(3)

For all the feedwater heaters in the process, the ΔTmin of 5 �Cwasselected by heuristics. From the known temperature values of thestreams inlets and outlets from the shell and tube sides of the heatexchanger, ΔTlm was evaluated using Eq. (4) [29]:

DTlm ¼ ðTi � toÞ � ðTo � tiÞln�Ti � toTo � ti

� (4)

For the condenser, the log-mean temperature difference wasevaluated using Eq. (5) as given [29]:

DTlm ¼ ðto � tiÞln�Tsat � tiTsat � to

� (5)

The overall heat transfer coefficient is a measure of the sum ofseveral individual resistances to heat transfer and can be estimatedfrom known values of Q, A and ΔTmin using Eq. (6).

U ¼ QADTlm

(6)

For the process under consideration, information about the heattransfer area of all the heat exchangers in the networkwas obtainedfrom the plant on visitation to the plant site for data gathering.Clarifications on this information were made where necessary bypersonal communicationwith the plant operators. The informationwas used for the estimation of the base case heat transfer coeffi-cients which formed the basis for simulation for sizing of heatexchanger network. The estimation of the heat exchanger heattransfer area for varying condenser cooling water flowrate was alsodone using Eq. (6).

The program codes are capable of evaluating the correspondingtemperature and enthalpy values of all the streams of material inand out of the shell and tube sides of the heat exchangers in thenetwork and estimating the new heat exchanger duty in thenetwork for varying condenser cooling water flowrate. The oper-ating data for the base case were used to validate the programcodes simulation results. Table 1 presents the actual plantmeasured data with their uncertainties and the calculated data forthe base case. Comparison of the two showed that the calculateddata were in agreement with the measured data. The programcodes are also capable of estimating the effect of cooling waterinlet temperature on plant performance at the base casecondenser cooling water flow rate and heat exchangers areas.

The plant operators’ experience from the operation of EgbinThermal Power Plant revealed that the maximum operating pres-sure limit for the condenser was 13 kPa at a temperature below100 �C after which it would automatically trip off. This piece ofinformation guided the selection of the range of condenser coolingwater flow rate for the thermal power plant simulation.

3.2. Heat exchanger costing

The relationship between equipment size and bare cost of heatexchanger was plotted on a logelog plot using literature data [29]and Microsoft Excel spreadsheet, and the best straight line fittedto the plot. The purchase costs of the heat exchangers that take intoaccount, the material of construction, operating pressure range andinflation rate were therefore obtained using Eq. (7).

Heat Exchanger Purchase CostðPCEÞ¼ Bare Cost From Cost Curve�Material Type Factors

� Pressure Factor� Index Factor

(7)

The index factor that relates the present year cost index to thepast year cost index was evaluated using Eq. (8):

Index FactorðI:FÞ ¼ Cost index for Present YearCost index for Past Year

(8)

The factorial method for capital cost estimation was used in theheat exchanger costing [29]. The physical plant cost that takes intoconsideration the equipment erection, piping, instrumentation andelectrical work is given in Eq. (9):

Physical Plant CostðPPCÞ ¼ PCE�1þ

Xfj�

(9)

In the estimation of equipment fixed capital cost, Eq. (10) thatincludes consideration for design and engineering, contractor’s feeand contingency was used.

Fixed Capital ¼ PPC�1þ

Xfi�

(10)

The fixed capital cost of the heat exchanger was amortized overthe useful life of the plant using Eq. (11) [30]:

Ar ¼ Inv

"jð1þ jÞN

ð1þ jÞN�1

#(11)

Table 2Response of condenser temperature and pressure to varying condenser coolingwater flowrate.

A.N. Anozie, O.J. Odejobi / Applied Thermal Engineering 31 (2011) 4083e4090 4087

A useful life period of 25 years and an annual interest rate of 12%were adopted. It was assumed that plant operates for 8000 h peryear.

Condenser cooling waterflowrate (m3/h)

32000 32200 32400 32600 32660(Base case)

Condenser Temperature (�C) 51 48 46 43 42Condenser Pressure (kPa) 12.3 11.5 10.0 8.7 8.3

3.3. Estimation of energy consumption, energy savings and plantefficiency

In the costing of energy, consideration was given to the type offuel used, thermal power plant generator output, the boiler effi-ciency and auxiliary power requirements for pumping. The fueltype considered was natural gas. The enthalpy of feedwater to theboiler could play a vital role in the fuel consumption in the furnacefor steam generation in the boiler. The more conserved the heatrejection from the cycle the more heat available within the cycle toimprove the quality of boiler feed water [26]. The unit heat rate,which is a measure of the quality of boiler feedwater and reheatsteam in relation to boiler efficiency and generator output wasdetermined using Eq. (12) [26]:

Unit Heat RateðkJ=kwhÞ

¼W1

�Hbo � Hfi

�þW2ðHro � HriÞ

Generator Output� Efficiency of Boiler Unit(12)

The rate of fuel consumption for steam generation whichdepends on the quality of feedwater and boiler efficiency wasestimated using Eq. (13) [26]:

Fuel Consumption Rateðkg=hÞ

¼ Unit Heat RateðkJ=kWhÞ � Generator OutputðkWÞCalorific value of fuelðkJ=kgÞ

(13)

The cost of fuel consumed was estimated using Eq. (14):

Cost of fuel consummedð$=yrÞ¼ Fuel consumption rate� Cost of fuel (14)

The percentage fuel saving was calculated by Eq. (15):

%Fuel Saving ¼ ðExisting plant fuel consumpion rate�Modified plant fuel consumpion rateÞExisting plant fuel consumpion rate

� 100 (15)

Invariably the cost of fuel saved was estimated using Eq. (16):

Cost of Fuel Savedð$=yrÞ ¼ Fuel saved� Cost of fuel (16)

In the overall estimation of energy cost it was assumed that theelectricity cost associated with pumping and other power auxiliaryconsumption in the plant was 10% of the generated power. Hencethe total energy cost was expressed mathematically as in Eq. (17):

Energy Cost ¼ Fuel Cost þ Cost of Electricity (17)

The total cost was therefore the sum of amortized fixed capitalcost and total energy cost.

Total Cost ¼ Fixed Capital Costþ Energy Cost (18)

The payback period is the length of time taken to recover themoney spent on an investment. It was calculated as:

Payback Period ¼ Fixed Capital CostCost Savings Achieved

(19)

The cycle efficiency is the measure of the system capacity toproduce the desired effect. In order to determine how well thedesired effect of the system was accomplished, the efficiency wascalculated as given in Eq. (20) [26]:

Cycle Efficiency�hcycle

�¼ Hbo � Hc

Hbo � Hbi(20)

4. Results and discussion

4.1. Plant performance at varying condenser cooling water flowrate

The response of condenser temperature and pressure to varyingcondenser cooling water flowrate at ambient temperature of 30 �Cis presented in Table 2. It was observed that condenser temperatureand pressure increased with decreasing condenser cooling waterflowrate. These observations are in agreement with that of Ga�nẚnet al. [27]. The constraint of maximum condenser pressure of13.0 kPa was used as a guide in setting the limit of feasiblecondenser cooling water flow rate in Egbin plant.

The fuel saved in a 220 MW unit and the enthalpy of boilerfeedwater at varying condenser cooling water flowrate is presentedin Fig. 3. There is a link between the quality of boiler feed water andthe fuel consumption in the furnace for raising the superheatedsteam in the boiler. The higher the quality of boiler feed water thelower the fuel consumption. Increasing condenser cooling waterflowrate reduced recovery of waste heat from the cycle. The

implication of this is seen in the energy cost. The dependence offuel cost and boiler feedwater enthalpy on condenser coolingwaterflowrate is shown in Fig. 4. It was observed that as the condensercooling water flowrate increasedmore cost was incurred in fuellingthe furnace to generate the superheated steam at the requiredtemperature and pressure.

The heat exchangers areas estimated for varying circulationwater flowrate is shown in Table 3. The total heat exchangers heattransfer areas increased with increase in circulationwater flowrate.It was observed that the condenser made the highest contributionin increasing the transfer areas requirement. The reason is that asthe circulation water increases more heat is rejected from the cyclethrough the condenser and consequentlymore heat transfer area inthe condenser will be required.

Fig. 5. Total cost and cycle efficiency versus circulating water flowrate.Fig. 3. Dependence of Fuel Saved on the Boiler Feed Water Enthalpy.

A.N. Anozie, O.J. Odejobi / Applied Thermal Engineering 31 (2011) 4083e40904088

4.2. Estimation of heat exchanger capital cost

The bare cost (Cb) of a given shell and tube heat exchanger ofarea (A), carbon shell, stainless steel tubes was derived as:

Cbcs ¼ 2089A0:808 (21)

The bare cost for stainless steel shell, stainless steel tubes heatexchanger was derived as:

Cbss ¼ 4410A0:699 (22)

The annualized capital cost for varying condenser cooling waterflowrate is also shown in Table 3. It was observed from Table 3 thatthere is direct relationship between the heat exchanger total heat

Fig. 4. Dependence of Fuel cost and Boiler feed water enthalpy on condenser coolingwater flowrate.

Table 3Estimation of heat exchanger area for varying condenser cooling water flowrate.

Condenser cooling water flowrate (m3/h) 32000 322

Equipment Area (m2) Are

Condenser 5401.13Air ejector 32.56Gland condenser 44.59Drain cooler 324.56LPH1 452.46LPH2 169.34LPH3 377.17HPH5 656.86HPH6 654.48Total Area 8113.16Condenser’s contribution to heat transfer area (%) 66.57Annualized capital cost ($/y) 12,271,064.3 13,

transfer areas and annualized capital cost. As the heat transfer areaincreases with increase in condenser cooling water flowrate, theannualized capital cost of equipment is also on the increase. This isdue to additional cost requirement for equipment modification tocope with the new heat transfer area.

4.3. Identification of optimum condenser cooling water flowrate

Two objectives indices, the annualized total cost of the plant andthe plant cycle efficiency were chosen for identifying optimumcondenser cooling water flowrate. The total cost estimated assummation of annualized capital cost and the energy cost wasobtained for varying condenser cooling water and is presented inFig. 5. On the secondary axis of the same Figure was plotted theplant cycle efficiency. The cycle efficiency ranged between 43 and45% which is in agreement with the results of first-law efficiency ofthermal plants from previous studies [1,6]. It was observed fromFig. 5 that the total cost increased with increase in condensercooling water flowrate while the plant cycle efficiency reduced. Itwas also observed that the flowrate that gave the minimum totalcost and maximum cycle efficiency was 32,000 m3/h.

4.4. Comparison of base case and optimum case performances

The cost analysis and cycle efficiency for the base case andoptimum condenser cooling water flowrate are shown in Table 4. Itcould be observed from the Table that the annualized capital costfor the optimum condition is lower than for the base case. This isdue to less heat transfer area required in the condenser at optimumflow rate. The percentage fuel saving estimated from the unit heatrate and steam rate for steam generation at optimum condensercooling water flowrate was 3.8% with the base case as reference.

00 32400 32600 32660

a (m2) Area (m2) Area (m2) Area (m2)

6431.61 7939.14 10374.56 10630.0031.72 30.94 30.20 30.0043.42 42.33 41.30 41.00

311.39 302.43 147.82 140.00457.24 462.09 491.62 500.00175.59 182.34 398.06 410.00384.44 387.48 388.20 390.00650.30 638.63 546.48 535.00649.80 643.20 580.07 580.00

9135.52 10628.56 12998.31 13256.0070.40 74.70 79.81 80.19

232,770.4 14,584,489.0 16,592,621.17 16,809,876.5

Table 4Comparison of energy and cost savings and cycle efficiency of base case andoptimum case.

Data Base case(32660 m3/h)

Optimum case(32000 m3/h)

Existing unit heat rate (kJ/kWh) 9808.36 9808.36Proposed unit heat rate (kJ/kWh) 9808.36 9438.28Calorific value of fuel (kJ/kg) 60.42 60.42Previous fuel rate for steam

generation (kg/h)35,711,836.26 e

Present fuel rate for steamgeneration (kg/h)

e 34,364,405.95

Fuel saved per annum (kg/yr) (0.00) 10,799,444,868.39% Fuel saving 0.0% 3.8%Density of natural gas (kg/m3) 635 635Price of natural gas ($/MMBTU) 4.35 4.35Fuel cost ($/y) 71,119,320.93 68,435,943.12Electricity cost ($/yr) 18,773,333.33 18,773,333.33Energy cost ($/yr) 89,892,654.27 87,209,276.45Amortized fixed capital cost ($/yr) 2,143,258.75 1,564,560.32Operating cost ($/yr) 92,035,913.01 88,773,836.78Fuel cost saving per unit ($/yr) e 2,683,377.81Fuel cost saving in 6 turbine units ($/yr) e 16,100,266.88Existing cycle efficiency 43.0 e

Proposed cycle efficiency e 45.0

Table 6Estimated Payback Period.

Description Amount

Fixed capital for plant modification ($) 4,694,220.96Fuel cost saved ($/yr) 2,683,377.81Payback period (year) 1.8

A.N. Anozie, O.J. Odejobi / Applied Thermal Engineering 31 (2011) 4083e4090 4089

The low energy cost at the optimum condition is due to conser-vation of waste heat in the cycle. From Table 4 the fuel cost savingper turbine unit for the optimum condenser cooling water flowratewas $2,683,377.81/yr and the saving possible for all the six unitswas $16,100,266.88/yr. There was also an increase in the plantcycle efficiency by 2%.

4.5. Cost implications of process improvement

The heat exchangers areas for heat transfer at the base case andoptimum case and the changes required to meet the optimum caserequirement and the cost implications are shown in Table 5. Itshowed that smaller condenser area was required at optimumcondition. Hence, there was no need for any modification as theexisting condenser area could cope with the present area require-ment. Also, the existing LPH2 and LPH3 areas could cope with thenewareas requirements and therewas no need of anymodification.The air ejector and gland condenser require new shells because ofsmall variations in the existing and the new area requirements.There were wide variations in the existing areas and the new areasrequirements for the drain cooler and HPH5 and HPH6 and theseexchangers needed to be replaced entirely with new ones. The costof changing exchanger shell was taken as half the cost of buyinga new exchanger. The total purchase cost of heat exchangers isshown in Table 5.

Table 5Heat exchanger cost estimates for optimum water circulation rate of 32,000 m3/h.

Heat exchanger Area (m2)(Base case)

Area (m2)(Opt. case)

Change required

Condenser 10,630.00 5401.13 Existing exchangerAir ejector 30.00 32.56 new shellGland condenser 41.00 44.59 new shellDrain cooler 140.00 324.56 New exchangerLPH1 500.00 452.46 New exchangerLPH2 410.00 169.34 Existing exchangerLPH3 390.00 377.17 Existing exchangerHPH5 535.00 656.86 New exchangerHPH6 580.00 654.48 New exchangerTOTAL (Purchase Cost)

P.F ¼ Pressure Factor; C. I ¼ Index Factor; M.F ¼ Material Type Factor

The total purchase cost of heat exchangers in Table 5 and Eqs.(9e11) were used to obtain the fixed capital for heat exchangersmodification and replacement shown in Table 6. The fuel costsaving achievable from plant modification is also presented in thesame Table. Using these values the payback period was obtained as1.8 years.

4.6. Sensitivity of plant at base case to varying condenser coolingwater inlet temperature

The response of the plant to variation in the condenser coolingwater inlet temperature at base case (cooling water flowrate of32,660 m3/h and total heat exchangers areas of 13256.0 m2) ispresented in Table 7. It was observed that the lower the condensercooling water temperature the higher the vacuum (lower pressure)in the condenser which is in agreement with the finding of Ga�nẚnet al. [27]. Furthermore, according to Ga�nẚn et al. [27], high vacuumin the condenser is required for high power generation from thelow pressure turbine. However, the power generated by the lowpressure turbine as a function of the condenser cooling watertemperature was not calculated in this study. Also the efficiencydefinition used by Ga�nẚn et al. [27] was based only on the lowpressure turbine and is different from the plant cycle efficiencydefinition used in this study.

It was observed in Table 7 that as the condenser cooling watertemperature increased, the condenser temperature and boiler feedwater enthalpy increased. At high cooling water temperature, thecondenser temperature is high and the condensate leaving thecondenser for the hot well exists at high temperature and hencecontributes to heat needed for preheating the boiler feedwater. Thecycle efficiency is a function of heat available within the cycle todrive the process. This is the reason for high cycle efficiency of 43%obtained at the high cooling water temperature of 30 �C when theboiler feed water enthalpy was 1021.80 kJ/kg.

Also shown in the Table 7 is the fuel consumption at differentcooling water inlet temperatures. The high availability of heatfor boiler feed water preheating at high cooling water inlettemperature is responsible for reduction in the fuel consumption inthe process as the cooling water temperature is increased. Thedecreasing trend in the fuel consumption with increasing trend inthe cycle efficiency is in agreement with results obtained byEskandari and Behzad [1].

Bare cost ($) P (kPa) C.I P.F M.F Purchasecost ($)

e 12.28 1.12 1.0 0.85 e

34,849 12.7 1.12 1.0 0.85 33,17744,925 13.48 1.12 1.0 0.85 42,769

223,392 17.76 1.12 1.0 0.85 212,669292,178 231.7 1.12 1.0 0.85 278,154

e 343.5 1.12 1.0 0.85 e

e 599.9 1.12 1.0 0.85 e

411,012 1761 1.12 1.1 0.85 430,411409,969 3087 1.12 1.3 0.85 507,378

1,416,326 1,504,558

Table 7Sensitivity of the plant at base case to varying condenser cooling water inlet temperature

Cooling Water Inlet Temperature (�C) 25 26 27 28 29 30Condenser Pressure (kPa) 6.94 7.20 7.48 7.76 8.04 8.30Condenser Temperature (�C) 35.2 36.5 38 39.5 42.1 42.1Boiler Feedwater Enthalpy (kJ/kg) 851.50 885.56 919.68 953.68 987.74 1021.80Fuel Consumption (kg/s) 10,514.04 10,399.65 10,281.00 10,162.23 10,043.71 9925.06Cycle Efficiency (%) 40.2 40.8 41.3 41.9 42.2 43.0

A.N. Anozie, O.J. Odejobi / Applied Thermal Engineering 31 (2011) 4083e40904090

5. Conclusion

The condenser circulation water flowrate that optimized thetotal cost of the investigated thermal power plant within thecondenser temperature and pressure tolerance limits was deter-mined. The improvement achieved on the plant cycle efficiency,heat transfer area requirements, annualized capital cost and fuelcost saving was estimated. It was found that modification of someplant components, especially the heat exchangers heat transferareas was required to cope with varying condenser cooling waterflowrate. The cost implication of the plant modification was eval-uated and the economic viability of the project was evaluated usingpayback period as the economic index. Sensitivity analysis of thethermal plant at the base case showed that the plant is sensitive tochanges in the condenser cooling water inlet temperature.

Acknowledgement

The authors gratefully acknowledge the assistance receivedfrom the Management of Egbin Thermal Power Plant, Ikorodu,Lagos State, Nigeria.

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Glossary

A: surface area of the heat exchangerAr: annuity [US$/year]CP: heat capacity flowrate of the fluid (kW/K)fi: factor for direct costs estimationfj: factor for indirect costs estimationH: enthalpy of the fluid (kJ/kg)Hbi: enthalpy of feedwater into the boiler(kJ/kg)Hbo: enthalpy of outlet steam from boiler(kJ/kg)Hc: enthalpy of steam into the condenser(kJ/kg)Hfi: enthalpy of the feedwater into the boiler (kJ/kg)Hi: enthalpy of stream i (kJ/kg)Hj: enthalpy of stream j (kJ/kg)Hri: enthalpy of the steam inlet into the reheater (kJ/kg)Hro: enthalpy of the reheated stream (kJ/kg)Inv: investment valuej: interest rateN: number of periods (years)n: number of mixing streamsQ: Exchanger duty (kW)T: temperature of the fluid in the shell side (�C)t: temperature of the fluid in the tube side (�C)Tsat: saturation temperature of the vapour condensate (�C)U: overall heat transfer coefficientW: mass flowrate of the fluid (kg/s)Wi: flowrate of stream i (kg/hr)W1: steam flowrate from boiler(kg/hr)W2: reheated steam flowrate(kg/hr)DTlm: log-mean temperature difference (�C)

Subscriptscs: carbon steelS: shellss: stainless steelT: tubei: inleto: outlet