Selection of Gas Turbine Air Bottoming Cycle for Polish compressor

Open Access Journal

Journal of Power Technologies 93 (2) (2013) 67–77journal homepage:papers.itc.pw.edu.pl

Selection of Gas Turbine Air Bottoming Cycle for Polish compressorstations

Daniel Czaja∗ , Tadeusz Chmielniak, Sebastian LepszyInstitute of Power Engineering and Turbomachinery, Silesian University of Technology

Konarksiego 18, Gliwice 44-100, Poland

Abstract

Gas turbines are one of the basic technologies used to produce electricity and power working machinery. Thepopularity of the technology results from its advantages, the most important of which are: fast start-up, highefficiency, low pollutant emissions, short construction time and reasonable size. Gas turbines are becomingincreasingly important in new power installations.

Combined cycles can be used to increase the efficiency of energy systems. Examples of high efficiencycycles are combined cycle power plants (CCPP) and gas turbine air bottoming cycles (GT-ABC), which area combination of a gas turbine and air turbine cycle coupled by means of a heat exchanger referred to as theAir Heat Exchanger (AHX). The main feature of the GT (and the GT-ABC) is low water consumption. Forthis reason, it can be used in gas transport and storage systems. The construction of this type of systems mayturn out to be energy-effective due to the advancement in flow machinery construction, especially in the fieldof improvement to blade profiles and sealing.

This paper presents an application of gas-air systems with example configurations. Three technologicalstructures are taken into consideration – a simple system ABC, an ABC with one air intercooler, and anadvanced cycle with two intercoolers. To improve the efficiency of air turbine installations, more complexsystem configurations are required.

Keywords: Air Bottoming Cycle, Gas Turbine Air Bottoming Cycle, Air Heat Exchanger, Gas turbine, Airturbine, Intercooler

1. Introduction

In a conventional combined cycle power plant(CCPP) the exhaust heat of the gas turbine (GT)is utilized by the heat recovery steam generator(HRSG). In addition to the boiler and steam turbine

∗Corresponding authorEmail addresses: [email protected] (Daniel

Czaja∗), [email protected] (TadeuszChmielniak), [email protected] (SebastianLepszy)

(ST), the cycle incorporates a condenser, pumps, pip-ing, heat exchangers and many other devices, result-ing in considerable costs. Importantly, even an ad-vanced steam bottoming cycle (SBC) provides onlyone third of total plant power. Another disadvantageof the SBC is the long start-up period. The use ofan air bottoming cycle (ABC) offers a significant in-crease in power and efficiency without the complex-ity of the SBC. ABC, which requires no water (nosteam), looks attractive for power generation in smallunits (under 50 MW range) [1] and in compressor

Journal of Power Technologies 93 (2) (2013) 67–77

stations. According to [2] there are two main objec-tives of ABC:

• to provide a combined thermodynamic cyclesystem that is coupled to a gas turbine andwhich has a greater thermodynamic efficiencythan the GT alone,

• to retain the operational flexibility and reliabil-ity of the GT,

The ABC thermal engine involves a multi-stage com-pression process where air passes through stages ofcompression and cooling. Prior to expanding in theturbines, the air passes through an air heat exchanger(AHX). An important component in terms of the ef-ficiency of the entire configuration of the gas turbinewith ABC is the AHX coupling the two systems. Themost important parameters impacting the optimumselection of the AHX are pressure drops and tem-perature differences. High efficiencies of the cycleare achieved for small temperature differences in theAHX, which result in its large dimensions. Otherpossible structural solutions are: shell-and-tube orplate heat exchangers. For modeling purposes, plateheat exchanger (PHE) was applied.

In the case of a gas turbine unit, the hot gasesare expanded in a turbine to produce torque, whichis used to drive an air compressor. The remainingtorque is available on an output shaft [3]. This can beused, for example, to drive a natural gas compressorin a gas station. The option of GT with steam reform-ing for exhaust heat recovery in a remote compressorstation applications was discussed by [4]. Configu-ration of GT with ABC was also analyzed by [5–7].Performance analyzes of gas turbine air – bottomingcombined system and industry applications analysiswere carried out by [8–10]. Thermodynamic anal-ysis of ABC as well as the results of a feasibilitystudy for using ABC for gas turbines waste heat re-covery/power generation on oil/gas platforms in theNorth Sea are to be found in [11].

Research on gas-air systems was done by [12]; itfocused on applications in heat engineering and sys-tems used to power working machinery in the gastransport and storage industry [13]. The use of gas-air systems as a heat source for the CCS installationof a coal-fired power unit was also done by [14].

Figure 1: Comparison between individual gas turbine systems

Thermodynamic analysis of ABC as well as exergyanalysis of GT-ABC was conducted by [15–17].

Gas-air systems can find application:

• in the food sector (industrial bakeries, powderedmilk factories),

• as a source of hot air in furnaces for glass melt-ing,

• in high-temperature furnace systems where pre-heated air comes from ABC,

• in offshore locations; gas-air systems can beused to improve the efficiency of simple powerunits with gas turbines operating at locationswithout access to large amounts of fresh water.

The mechanical power obtained from the turbine canbe used either to support the gas turbine system or togenerate electricity. Owing to the short start-up timeof the air turbine, the ability to meet peak demandfor power may also be significant. Air turbine sys-tems are simple in terms of operation. This is due tothe absence of: (i) a combustion process, (ii) toxicmedia or (iii) media causing erosion or needing to betopped up. ABC also improves the performance ofthe new combined engine in off-design mode [2]. Itis expected that ABC will require much lower invest-ment expenditure in respect of CCPP [18].

Fig. 1 shows a comparison in a T-s diagram be-tween: recuperated gas turbine, gas turbine air bot-toming cycle (GT-ABC) – also known as dual gasturbine – and gas turbine steam bottoming cycle (GT-SBC).

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Table 1: Włocławek compressor station basic data [19, 20]

Parameter Value

Gas composition 96.3% CH4,2.89% C2H6

Gas inlet pressure (tocompressor station)

approx.6.05 MPa

Gas outlet pressure (fromcompressor station)

approx.8.35 MPa

Volume flow 7,694,000 Nm3/hGas inlet temperature(average annual)

9◦C

Gas outlet temperature <30◦C

Table 2: GT10 gas turbine basic data [19, 20]

Parameter Value

Shaft power 25 MWEfficiency (nominal point) 35.2%Revolutions 7700 rpmMax. revs of turbine 8625 rpm

The main goal of this paper is to determine the ba-sic thermodynamic characteristics of the systems un-der analysis. Increases in mechanical power outputand energy efficiency are determined. Improvementof the GT engine can be interesting from the point ofview of obtaining additional power.

2. Compressor station description

There are five gas compressor stations in thePolish section of the “Yamal Western Europe” gaspipeline. They are tasked with providing sufficientpressure to compensate for flow friction losses. Eachcompressor aggregate has 25 MW of power and con-sists of a centrifugal compressor and gas turbine

Table 3: 50P2 compressor basic data [18, 19]

Parameter Value

Suction pressure 6.4 MPaDischarge pressure 8.45 MPaRevolutions 6680 rpmFlow c. 2 million Nm3/h

unit. Compressor aggregates are grouped in par-allel. They also contain the gas coolers and theirown compressor and turbine control system. Ev-ery unit can be controlled from the Distributed Con-trol System (DCS). Nominal discharge pressure is8.45 MPa. Yearly, 40 billion Nm3 of gas can be com-pressed [19]. The basic data of an example compres-sor station are shown in Table 1. Technical data ofthe GT10 gas turbine and 50P2 compressor are pre-sented in Tables 2–3 [19, 20]. Nm3 refers to gas vol-ume at normal temperature and pressure, which are0◦C and 101.325 kPa.

This section of the 4,107 km-long Yamal – Eu-rope pipeline includes five compressor stations atCiechanow, Szamotuly, Zambrow, Wloclawek andKondratki. The stations have a total driving powerof 600 MW and feature four to six turbocompressorunits driven by a 25 MW gas turbine.

3. Technological structures under analysis

ABC is a Brayton cycle that utilizes the exhaustheat from a topping GT engine. The configurationwhere the ABC cycle recuperates heat from toppingGT cycles is also referred to as a dual gas turbinecombined cycle [21].

In a simple gas-air system a compressor with nointercooler (Case A), a heat exchanger AHX cou-pling the gas and air systems, and an air turbine areused. Referring to Fig. 2 a gas turbine includes acompressor and combustion chamber. The productsof combustion and excess air are applied to a tur-bine where they power an output shaft (and drive acompressor). After expansion, gases exit the turbineat high temperature. The exhaust gas duct conveysthem to an air heat exchanger where they give up asubstantial portion of their energy to heat up com-pressed air. The next step is to expand the air in theair turbine and produce torque.

To improve the efficiency of air turbine installa-tions, it is necessary to employ more complex sys-tem configurations. An example of a complex airturbine system is the installation shown in Fig. 3(Case B). A first compressor stage compresses am-bient air. Then compressed air is conveyed to anintercooler (IC) and cooled to the lowest possibletemperature. After intercooling the compressed air

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Figure 2: Simple gas-air system (Case A)

is again compressed in a second air compressor andthen air is heated up in AHX. Afterwards the air isexpanded in the air turbine as in Case A. The com-pression process is nearly adiabatic. The intercooleris used between an adjacent pair of compressors tokeep the temperature of the compressed air low be-fore compression and make the work of compressionmore efficient. The compressed air is passed throughAHX where its temperature is raised by the hot GT’sexhaust gas flowing in the AHX counterflow. After-wards the hot compressed air is expanded in the airturbine and torque is produced. Intercooling betweenthe compressors helps to increase the heat transfer inthe recuperator by reducing the overall compressordischarge temperature. The compressor work is alsoreduced, while the available work from the turbine isincreased.

The most complex installation considered is pre-sented in Fig. 4 (Case C). The outlet air from com-pressor C3 (system with one intercooler) and C4 isat lower temperatures than the simple system, whichallows for a more intense cooling of the gas turbineflue gases and a reduction in the driving operationof the compressors. It should be noted that the in-

Figure 3: Complex gas-air system with an air intercooler(Case B)

tercoolers here are air cooled heat exchangers. Thecompressor work is also reduced, while the availablework from the turbine is increased.

4. Calculations

The target of a GT-ABC system analysis should beto maximize the ratio of generated electricity to thechemical energy of the fuel. In this case, the energyefficiency of the system can be defined as follows:

ηelGT−ABC =NelGT + NelAT

m f · LHV(1)

ABC system efficiency can be evaluated using thefollowing definition of energy efficiency:

ηeAT =NmAT

Q4(2)

where: Q4 is the heat of cooling flue gases to thereference temperature.

In order to simplify the analyses, it is assumed inthe calculations presented in this paper that Q4 cor-responds to the cooling of flue gases to temperaturet5 = 15◦C, while the water vapor contained in fluegases is not condensed.

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Figure 4: Complex gas-air system with two air intercoolers (Case C)

By adopting some simplifying assumptions, it ispossible to determine the dependence between theenergy efficiency of the gas turbine and air turbinesystems, as well as of the entire system. The mostimportant simplifying assumption is that there are noradiation heat losses from either the gas or the airturbine. The dependence can be defined as follows:

ηe = ηeGT + ηeAT (1 − ηeGT ) (3)

Calculations of system parameters were per-formed using the authors’ own algorithms (ideal gasmodel; heat capacity as a function of gas tempera-ture).

The GT-ABC system is optimized in terms of en-ergy efficiency. The pressure ratio and the air massflow are chosen as decision variables. The othervalues are selected according to the parameters ofthe machines and equipment operating in gas turbinesystems. In Case A for set parameters of the heat ex-changer (AHX effectiveness), pressure value p2a isvaried within the determined range. The procedureis repeated for different mass flow values of the airsucked in by compressor C2.

In complex systems it is important to choose anappropriate pressure value at the outlet of each stageof the compressor. For example, in the case C theoptimum pressure value which minimizes the powerconsumption used to drive the compressor can bedetermined by minimizing the following objectivefunction [22].

[T1a (ς1 − 1) + T3a (ς3 − 1) + T5a (ς5 − 1)]1ηi→ min

(4)

where: ς1 = T2asT1a

, ς3 = T4asT3a

, ς5 =T6asT5a

The procedure is repeated for different values ofthe air mass flow sucked by the first stage compres-sor. The result where energy efficiency reaches thehighest value is taken as the optimized structure.

The pressure drops in the heat exchanger are theparameters which have a significant impact on the ef-ficiency of the entire system. The higher the pressuredrop in AHX, the lower the efficiency ηeAT . On theother hand, a bigger pressure drop makes it possibleto obtain high heat transfer coefficients, which leadsto a reduction in the heat exchange area and, conse-quently, to a smaller sized device. A plate heat ex-changer (PHE) was applied for modeling purposes.PHE are widely employed in chemical, food andpharmaceutical process industries. They are com-pact, easy to clean, efficient and very flexible [23]. Inorder to determine the heat transfer coefficient manyphysical properties of air and exhaust gas were takeninto account:

α = f(m, λ, cp, ρ, ν,∆t, a...

)(5)

The calculation algorithm is based on the LMTDmethod [24]. To determine the heat surface area theNusselt number should be found. The Nusselt num-ber for this type of heat exchangers can be defined bythe formula [25]:

Nu = 0.022 ·√ξ0 · β · βt · Re0.825 · Pr0.54 (6)

where: ξ0 – flow resistance ratio, β – turbulencedamping ratio, βt – forced turbulence ratio.

It is assumed that the wall which separates mediain the heat exchanger is made of P235GH and 16Mo3

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Figure 5: Energy efficiency as a function of pressure ratio(Case A)

steel, and its heat conductivity depends on tempera-ture. Also 13CrMo4-5 is taken into account for sur-faces which work in high temperature areas.

Depending on minimal product m · cp of consid-ered fluids, the effectiveness of AHX and IC can bedefined as follows:

ε =Tinhot − Touthot

Tinhot − Tincold(7)

if(m · cp

)min

=(m · cp

)cold

,

ε =Toutcold − Tincold

Tinhot − Tincold(8)

if(m · cp

)min

=(m · cp

)hot

5. Results and discussion

The results of energy efficiency and increase inmechanical power output for different pressure dropsin the AHX are presented in Fig. 5 and Fig. 6(Case A). The value of pressure drop always refers tothe inlet pressure of individual heat exchangers. Pre-sented results are obtained for polytropic efficiencyof turbomachinery ηpC2 = ηpT2 = 88% and effec-tiveness of AHX εAHX = 96%. The optimal air massflow is equal ma = 80.46 kg/s. For a pressure ratio

Figure 6: Mechanical power output as a function of pressureratio (Case A)

βC2 = P2/P1 = 3.88 (∆PAHX = 0%) the highest en-ergy efficiencies and the maximum value of mechan-ical power output are recorded. Efficiency is linearlydependent on mechanical power output. Therefore,the maximum increase in mechanical power outputand energy efficiency is for the same pressure ratio.The exhaust gas temperature from ABB GT10 valueis equal T4 = 813 K. A drop in the exhaust gas tem-perature results in a considerable decrease in energyefficiency.

Increase in mechanical power output was deter-mined according to formula 9.

∆N =NABC

NGT· 100% (9)

More significant increases in efficiency comparedto Case A are achieved only for high values of poly-tropic efficiency of turbo machines. Presented resultsfor Case B (Fig. 7–8) are obtained for polytropic ef-ficiency of turbomachinery ηpCi = ηpT2 = 88% andeffectiveness of AHX εAHX = 96%, εIC = 90%.The optimal air mass flow is almost the same likein case A (ma = 80.56 kg/s). The pressure ratio ofcompressor C2 is βC2 = P2/P1 = 2.41 and outletpressure of compressor C3 varied within the deter-mined range.

The energy efficiency obtained for various pres-

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Figure 7: Energy efficiency of a complex system as a functionof pressure ratio (Case B)

sure ratios was dependent on the technical structure.The cycle with two intercoolers had the highest pres-sure ratio (Fig. 9–10). This system also has high en-ergy efficiency across a wide range of pressure ratios.The pressure drop in heat exchangers shows a similartrend. The calculation made for polytropic efficiencywas ηpCi = ηpT = 0.88 and for effectiveness of AHXεAHX = 0.96 (εICi = 0.9 respectively).

Energy efficiency as a function of compressor ra-tio for each of analyzed cases is shown in Fig. 11.The curves illustrate two different levels of pressuredrop. The system with two intercoolers enjoys highenergy efficiency across a wide range of pressure ra-tios. On the other hand there is a very significantincrease in energy efficiency between GT-ABC withno inter-stage cooler and GT-ABC with one IC.

The dependence between AHX and intercooler ef-fectiveness and their influence on the energy effi-ciency of the cycle is presented in Fig. 8. Calcula-tions were made parametrically for the purpose ofcomparing the energy efficiency value of the refer-ence case (ηpCi = ηpT2 = 88%, εIC = 90%, εAHX =

96%, ηeGT−ABC = 46.28%).

Figure 8: Mechanical power output of a complex system as afunction of pressure ratio (Case B)

Figure 13: The impact of relative heat transfer surface area onenergy efficiency and AHX effectiveness

The heat exchanger AHX is characterized by alarge heat transfer surface area to achieve high heatexchanger effectiveness. The influence of heat trans-fer surface area on effectiveness and energy effi-

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Figure 9: Energy efficiency of a complex system as a functionof pressure ratio (Case C)

ciency was considered and the results are shown inFig. 12. The value 100% of relative heat transfer sur-face area represents a heat transfer surface area equalto about AAHX(REF) = 89, 000 m2 (for εAHX = 96%).The curves illustrate polytropic efficiency of turbo-machinery ηpC2 = ηpT2 = 88%. Increasing the heattransfer surface area results in a much larger AHX,but also higher energy efficiency of the system. Theimpact of relative heat transfer surface area on en-ergy efficiency and AHX effectiveness is presentedin Fig. 13 (results for Case A).

Table 4: Comparison between Cases A, B and C

Parameter Case A Case B Case C

∆PAHX = ∆PIC,%

4 4 4

ηeGT−ABC, % 42.61 43.71 44.03∆N, % 20.28 23.51 24.17βC2, - 3.88 2.12 1.92βC3, - N/A 2.78 2.04βC4, - N/A N/A 2.37AAHX, m2 ∼27,500 ∼27,000 ∼26,500AIC1, m2 N/A ∼10,600 ∼10,600AIC2, m2 N/A N/A ∼10,600ma, kg/s 80.92 81.06 81.08

Figure 10: Mechanical power output of a complex system as afunction of pressure ratio (Case C)

A comparison of Case A and Case B is presentedin Table 4. Calculations were made for polytropicefficiency of turbomachinery ηpCi = ηpT2 = 88%,effectiveness of intercooler εIC = 75%, effective-ness of AHX εAHX = 88% and the same pressuredrop value on the both sides of heat exchangers equal∆PAHX = ∆PIC = 4%. Optimal mass flow, pressureratio, increase in mechanical power output and heattransfer surface area were determined.

6. Conclusion

The analyzes presented in this paper show that theconsidered gas turbine systems coupled with an airsystem (GT ABC) is very interesting from the pointof view of power engineering. Their energy effi-ciency is higher than that of stand-alone gas turbineunits. The systems are characterized by high valuesof overall efficiency. The most efficient consideredinstallation of Case A can reach ηeGT−ABC = 44.31%when ηpC2 = ηpT2 = 88%, εAHX = 96% and pres-sure drop in AHX is omitted. The increase in me-chanical power output is ∆N = 24.95% for thiscase. Using an additional facility such as an in-tercooler (Case B), electricity generation efficiencyrose to ηeGT−ABC = 46.28% (∆N = 30.56%, ηpCi =

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Figure 11: Energy efficiency of analyzed cases

ηpT2 = 88%, εAHX = 96%, εIC = 90% and pres-sure drop is omitted). The most significant increasein energy efficiency and mechanical power outputis assigned to Case C, where two intercoolers wereadopted ηeGT−ABC = 47.12% (∆N = 32.63%, ηpCi =

ηpT2 = 88%, εAHX = 96%, εIC = 90% and pres-sure drop is omitted). The use of this type of in-stallation may take place if the economic analysisis strongly affected by the main advantages of GT-ABC systems, such as the low water consumption,operation flexibility and the potential to meet peakdemand for energy. The thermodynamic potential ofGT-ABC was shown. The energy efficiency and in-crease in mechanical power output were determined.A competitive cycle in respect of SBC for power en-gineering where is a lack of water was presented.

Further investigations are required to show all ad-vantages of the GT-ABC engine in relation to othercycles such as [26–28]. Improvement of existingengineering achieves higher values of energy effi-ciency and increases in mechanical power output onthe shaft (shafts). Further investigation may bring

Figure 12: The influence of AHX and IC effectiveness on en-ergy efficiency

to light additional improvements to the GT-ABC en-gine.

AcknowledgmentsThe results presented in this paper were obtained

from research work co-financed by the National Cen-tre of Research and Development in the frameworkof Contract SP/E/1/67484/10 – “Strategic ResearchProgramme – Advanced Technologies for obtainingenergy: Development of a technology for highly effi-cient zero-emission coal-fired Power units integratedwith CO2 capture”.

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Nomenclature

α Heat transfer coefficient, W/(m2K)

β Pressure ratio, -

∆ Difference, increase, -

ε Effectiveness,

λ Thermal conductivity, W/(m K)

ρ Density, kg/m3

υ Kinematic viscosity, m2/s

ζ Temperature ratio, -

m Mass flow rate, kg/s

cp Specific heat, J/(kg K)

A Heat transfer surface area, m2

a Thermal diffusivity, m2/s

AT Air Turbine

CMB Combustion chamber

G Generator

LHV Lover heat value, J/kg

m Dynamic viscosity, Pa·s

N Power, %

Nu Nusselt number, -

p,P Pressure, MPa

Pr Prandtl number, -

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R Gas constant, J/(kg K)

Re Reynolds number, -

T, t Temperature, K, ◦C

1,1a,2... refers to individual reference points (Fig.2-4)

a Air

ABC Air Bottoming Cycle

AT Air turbine

C Compressor

c Cycle

cold Cold medium

e Energetic

el Electric

f Fuel

G Generator

GT Gas turbine

hot Hot medium

i Internal, isentropic

in Inlet

m Mechanical

out Outlet

p Polytropic

ref Reference case

T Turbine

t Temperature, K, ◦C

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Selection of Gas Turbine Air Bottoming Cycle for Polish compressor

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