Máquina Térmica Ideal de Ciclo Brayton

Máquinas Térmicas Trabalho 1 bimestre Prof Antônio Salvador Neto 1 Introdução O presente documento visa orientar a atividade de avaliativa processual da disciplina de Máquinas Térmicas oferecida no 1 período de 2024 pela Multivix Vila Velha A atividade consiste no projeto de uma máquina térmica ideal de ciclo Brayton 2 Formação dos grupos Os alunos devem organizar grupos de 3 a 5 integrantes escolher um dos artigos relacionados no apêndice e informar ao professor em tempo de aula para que seja registrado Não será admitido mais de um grupo com o mesmo artigo A prioridade de escolha se dará por ordem de registro dos grupos 3 Entrega O trabalho deve ser entregue até o dia 17042022 por via eletrônica através do portal httpsavapmultivixedubrloginindexphp 4 Atividade A atividade consiste em realizar os cálculos de um ciclo Brayton ideal à ar que opera segundo algumas condições especificadas no anexo I Etapa 1 Parâmetro de projeto Cada aluno tem um conjunto de dados que são os parâmetros do projeto do ciclo Brayton no apêndice I O grupo deve realizar o trabalho com os dados do aluno que aparece primeiro na lista deste apêndice Etapa 2 Determinação da razão de pressão Os parâmetros de projeto não informam a razão de pressão Esse valor deve ser decidido pelo grupo Esse parâmetro costuma estar entre 2 e 12 mas não há restrição quanto à escolha O trabalho deve justificar a escolha desse valor Etapa 3 Estados termodinâmicos Calcule os estados termodinâmicos entre as transformações Etapa 4 Trabalho e calor Calcule o trabalho realizado pelo ciclo e os calores absorvido e rejeitado Etapa 5 Rendimento do ciclo Calcule a eficiência térmica do ciclo usando os resultados obtidos na etapa 4 e compare com o resultado se utilizando da equação de eficiência do ciclo Brayton em função da razão de pressão e da razão de calores específicos Etapa 6 Rendimento de Carnot Calcule o rendimento de Carnot para uma máquina térmica de Carnot operando sob as mesmas condições e disserte uma comparação entre os rendimentos obtidos 5 Avaliação O trabalho terá valor máximo de três pontos e mínimo de zero pontos e será avaliado baseado em critérios Cada critério representa uma parcela dos pontos do trabalho A avaliação de cada critério pode conferir um valor igual ou inferior ao valor da parcela que ele representa Os critérios de avaliação serão os seguintes Critério Ponto s O trabalho tem capa está organizado coerente usa linguagem padrão e descreve claramente o que se pede para ser feito na etapa 1 05 Descreve de forma clara e completa o que se pede na etapa 2 05 Descreve de forma clara e completa o que se pede na etapa 3 05 Descreve de forma clara e completa o que se pede na etapa 4 05 Descreve de forma clara e completa o que se pede na etapa 5 05 Descreve de forma clara e completa o que se pede na etapa 6 05 Apêndice I Aluno Pressão externa do ar kPa Temperatura externa do ar C Temperatura da combustão C Alex Araujo Minini 130 30 465 Breno Barbosa Duarte 88 25 423 Breno Martins Lordes Machado 103 13 474 Breno Souza Bonadiman Garcia 86 19 451 bruno alcantara dos santos 108 3 373 Caio Ferreira Lopes 100 42 466 Davi Lucas da Costa Sousa 91 10 420 Fabrycio Henrique Oliveira de Souza 145 41 405 Felipe Vassoler Dos Santos 115 3 412 Gabriel Antonio de Oliveira Souza 55 28 513 Gabriel Guzzo De Carvalho 90 21 485 Guilherme Amaral Gomes Coutinho Mata 121 24 520 Guilherme Loss Binda 98 22 425 Gustavo Dela Costa Freire 97 24 496 Ian Auer Guss 90 2 470 Jader Fernandes de Oliveira 78 15 500 João Felipe Tellau Vago 100 31 566 João Paulo Fagundes de Oliveira 82 39 445 Joao Pedro dos Santos 135 19 366 Julia Soares de Freitas 100 15 422 Lívia França dos Santos 73 6 360 Luana Aparecida dos Santos Bispo 103 27 457 Lucas Dos Santos Lucio 144 14 371 Luiz Filipe Lopes Campos Alves 70 10 475 Luiz Guilherme Campos Costa 116 2 401 Luiz Rycardo Rebuli Mesquita 112 16 516 Marco Aurélio Bezerra Silva 132 21 341 Mateus de Souza Santos Nascimento 142 38 398 Mateus Lahas Pazolini 114 23 466 Matheus Lima do Nascimento 56 27 484 Mauro Marcos Alves Pereira Junior 80 36 450 Pedro Henrique Finco dos Santos 110 25 434 Raphael Eduardo laiola rapozo 99 16 504 Rene Coli da Silva 113 14 364 Rodrigo Dantas de Souza santos 110 0 400 Rubson Luiz Trancoso Junior 114 48 437 Sabino Bispo Filho 117 15 525 Sara Gambarini do Carmo 91 5 425 Taillon Dias Costa 99 48 373 Thiago do Nascimento Pereira 57 2 457 Victor Bigossi de Camargos Pereira 125 28 355 Vinícius Alves 86 5 360 Vinicius Brazil da Penha 112 27 525 Vitor da Silva Vieira 94 0 367 Yago Firme dos Santos 85 1 505 Yasmim de Santana Cravo Constante 97 39 429 Trabalho Máquinas Térmicas Ciclo Brayton Etapa 1 e 2 Parâmetros de Projeto e Determinação da Razão de Pressão Conforme Apêndice I presente no trabalho do 1º Bimestre da disciplina de Máquinas Térmicas este grupo possui os seguintes parâmetros de projetos referente ao aluno Bruno Alcântara dos Santos Pressão Externa do Ar P1 130 kPa Temperatura Externa do Ar T 1 3ºC ou 276 K Temperatura de Combustão T 3 373ºC ou 646 K Para selecionar a razão de pressão da turbina o grupo se baseou na turbina a gás da General Eletri c PG5371 cuja razão de pressão r p é de 106 sendo utilizada para a geração de energia httpswwwgevernovacom A escolha deste modelo em relação aos demais se deve ao mesmo possuir a menor temperatura de exaustão ou seja o que se aproximaria mais a temperatura de combustão do projeto do grupo Na tabela abaixo é possível observar os diferentes modelos e as suas principais características sendo o circulado de vermelho o modelo utilizado como referência neste trabalho Tabela 1 Modelos de Turbinas a Gás General Eletri c Fonte httpswwwgevernovacom Etapa 3 Estados Termodinâmicos Para o cálculo da Etapa 3 cada estado será calculado separadamente obtendose sempre os parâmetros temperatura pressão e entalpia Para isso será utilizado o modelo de ciclo padrão a ar e a Tabela A7 do Apêndice A da 8ª Edição do Livro Fundamentos da Termodinâmica de Bornakke Sonntag 2018 Estado 1 Dados fornecidos Pressão Externa do Ar P1 130 kPa Temperatura Externa do Ar T 1 3ºC ou 276 K Da Tabela A7 interpolando entre os valores de temperatura de 260 e 280 K temse que h127638 kJkg Estado 2 Como temse um processo de compressão isentrópica de 1 para a 2 podese afirmar que T2 T1 P2 P1 k 1 k T 2T 1r p k1 k Considerando o fluido de trabalho como o ar temse que k14 assim T 2276 106 0 4 1 454182K Assim como P2 P1 r p P2r pP11061301378kPa Da Tabela A7 interpolando entre os valores de temperatura de 540 e 560 K temse que h254658kJ kg Estado 3 Dados fornecidos Temperatura de Combustão T 3 373ºC ou 646 K No ciclo Brayton o processo de combustão é considerado como o processo de fornecimento de calor à pressão constante assim podemos afirmar que P3P21378kPa Da Tabela A7 interpolando entre os valores de temperatura de 640 e 660 K temse que h36559kJ kg Estado 4 No ciclo Brayton o processo de 3 para 4 é um processo de expansão isentrópica devendo a pressão voltar as condições iniciais portanto a razão de pressão neste processo será o inverso do que ocorre no processo de compressão ou seja P4P31106 assim P4 1 106 P3130kPa T 4 T 3 P4 P3 k1 k T2T 1 P4 P3 k 1 k T 4646 1 106 0 4 1 432907 K Da Tabela A7 interpolando entre os valores de temperatura de 320 e 340 K temse que h43297kJ kg Na Tabela 2 temse um resumo dos parâmetros obtidos Tabela 2 Estados Termodinâmicos do Ciclo Estados Temperatura K Pressão kPa Entalpia h 1 276 130 27638 2 54182 1378 54658 3 646 1378 6559 4 32907 130 3297 Fonte Próprio Autor Etapa 4 Trabalho e Calor Para facilitar na obtenção do trabalho e calor o grupo considerou o Ciclo Brayton apresentado na Figura 1 no qual temse um trabalho específico consumido no compressor wc um trabalho específico produzido na turbina wt um calor absorvido na câmara de combustão trocador de calor entre o compressor e a turbina qh e um calor rejeitado no trocador de calor entre a turbina e o compressor ql Figura 1 Ciclo Brayton Fonte Bornakke Sonntag 2018 Analisando pela 1ª Lei o compressor e considerandoo adiabático e atuando em regime permanente temse que d EV C dt Q W menthe msai hs 00 W c m h1 mh2 wch2h154658276 38 wc2702kJ kg Analisando pela 1ª Lei a câmara de combustão e considerandoo que atua em regime permanente temse que d EV C dt Q W menthe msai hs 0 Qh m h2 m h3 qhh3h2655954558 qh110 32kJ kg Analisando pela 1ª Lei a turbina e considerandoa adiabática e que atua em regime permanente temse que d EV C dt Q W menthe msai hs 00 W t m h3 mh4 wth3h4655 93297 wt3262kJ kg Analisando pela 1ª Lei o trocador de calor responsável pela rejeição de calor considerando que o mesmo atua em regime permanente temse que d EV C dt Q W menthe msai hs 0 Ql m h4 mh1 qlh4h1329727638 ql5332kJ kg Etapa 5 Rendimento do Ciclo Da definição de rendimento para máquina térmica temse que η1wliq qh qhql qh 1 ql qh 1 5332 11032 η105167ou5167 Para o modelo de ciclo padrão a ar frio no qual os calores específicos são constantes o rendimento térmico do ciclo depende exclusivamente da razão de pressão e da razão de calores específicos do ar k assim sendo η21 1 r p k1 k 1 1 106 04 1 4 η204906ou496 Comparando os rendimentos η1eη2 notase que o rendimento do ciclo utilizando o modelo de ciclo padrão a ar η1 possui um rendimento maior que o do modelo de ciclo padrão a ar frio η2 isso se deve ao modelo de ciclo padrão a ar considerar que haja variação nos calores específicos conforme há um aumento da temperatura efeito que proporciona um aumento da eficiência do ciclo Etapa 5 Rendimento de Carnot Considerando que o Ciclo Brayton estudado fosse uma máquina térmica de Carnot o rendimento térmico seria o máximo possível uma vez que a máquina térmica não possui irreversibilidades Assim sendo a temperatura da fonte quente seria a temperatura de combustão e a temperatura da fonte fria seria a temperatura do ar externo Com isso o rendimento de Carnot seria de ηcarnot1T l T h 1276 646 ηcarnot05758ou5758 Isso demonstra que o Ciclo Brayton estudado possui uma eficiência menor que a de Carnot e portanto poderia ser teoricamente obtido Entretanto notase claramente que as eficiências são próximas desprezando efeitos claramente existentes como a transferência de calor da turbina e compressor para o ambiente e a perda de pressão nos trocadores de calor concluindo que este Ciclo Brayton é apenas uma idealização e não corresponderia ao que aconteceria na prática se o mesmo ciclo fosse projetado e construído GE Power Systems GE Gas Turbine Performance Characteristics Frank J Brooks GE Power Systems Schenectady NY GER3567H Float Plane Fun The Northwest Experience Pacific Coastal Airline World Famous Experience one of the most beautiful sights in Canada aboard one of our exciting daily scenic float plane tours This 45minute flight is packed with fun and fantastic views of our mountain ranges including Mount Garibaldi Golden Ears Mount Baker Britannia Mines Squamish and where the Pacific Ocean meets the mighty Fraser River Tours operate daily yearround from our terminal at Vancouver Harbour on Coal Harbour Contents Introduction 1 Thermodynamic Principles 2 The Brayton Cycle 3 Thermodynamic Analysis 6 Combined Cycle 7 Factors Affecting Gas Turbine Performance 8 Air Temperature and Site Elevation 8 Humidity 8 Inlet and Exhaust Losses 9 Fuels 10 Fuel Heating 11 Diluent Injection 12 Air Extraction 12 Performance Enhancements 12 Inlet Cooling 13 Steam and Water Injection for Power Augmentation 14 Peak Rating 14 Performance Degradation 14 Verifying Gas Turbine Performance 15 Summary 15 List of Figures 16 List of Tables 16 GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 i GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 ii Introduction GE offers both heavyduty and aircraftderivative gas turbines for power generation and industri al applications The heavyduty product line con sists of five different model series MS3002 MS5000 MS6001 MS7001 and MS9001 The MS5000 is designed in both single and twoshaft configurations for both generator and mechanicaldrive applications The MS5000 and MS6001 are geardriven units that can be applied in 50 Hz and 60 Hz markets All units larger than the Frame 6 are direct drive units The MS7000 series units that are used for 60 Hz applications have rotational speeds of 3600 rpm The MS9000 series units used for 50 Hz applications have a rotational speed of 3000 rpm In generatordrive applica tions the product line covers a range from approximately 35800 hp to 345600 hp 26000 kW to 255600 kW Table 1 provides a complete listing of the avail able outputs and heat rates of the GE heavyduty gas turbines Table 2 lists the ratings of mechani caldrive units which range from 14520 hp to 108990 hp 10828 kW to 80685 kW The complete model number designation for each heavyduty product line machine is pro vided in both Tables 1 and 2 An explanation of the model number is given in Figure 1 This paper reviews some of the basic thermo dynamic principles of gas turbine operation and explains some of the factors that affect its performance GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 1 Table 1 GE gas turbine performance characteristics Generator drive gas turbine ratings GE Generator Drive Product Line Model Fuel ISO Base Heat Heat Exhaust Exhaust Exhaust Exhaust Pressure Rating Rate Rate Flow Flow Temp Temp Ratio kW BtukWh kJkWh lbhr kghr degrees F degrees C x103 x103 PG5371 PA Gas 26070 12060 12721 985 446 905 485 106 Dist 25570 12180 12847 998 448 906 486 106 PG6581 B Gas 42100 10640 11223 1158 525 1010 543 122 Dist 41160 10730 11318 1161 526 1011 544 121 PG6101 FA Gas 69430 10040 10526 1638 742 1101 594 146 Dist 74090 10680 10527 1704 772 1079 582 150 PG7121 EA Gas 84360 10480 11054 2361 1070 998 536 127 Dist 87220 10950 11550 2413 1093 993 537 129 PG7241 FA Gas 171700 9360 9873 3543 1605 1119 604 157 Dist 183800 9965 10511 3691 1672 1095 591 162 PG7251 FB Gas 184400 9245 9752 3561 1613 1154 623 184 Dist 177700 9975 10522 3703 1677 1057 569 187 PG9171 E Gas 122500 10140 10696 3275 1484 1009 543 126 Dist 127300 10620 11202 3355 1520 1003 539 129 PG9231 EC Gas 169200 9770 10305 4131 1871 1034 557 144 Dist 179800 10360 10928 4291 1944 1017 547 148 PG9351 FA Gas 255600 9250 9757 5118 2318 1127 608 153 Dist 268000 9920 10464 5337 2418 1106 597 158 GT22043E Thermodynamic Principles A schematic diagram for a simplecycle single shaft gas turbine is shown in Figure 2 Air enters the axial flow compressor at point 1 at ambient conditions Since these conditions vary from day to day and from location to location it is convenient to consider some standard condi tions for comparative purposes The standard conditions used by the gas turbine industry are 59 F15 C 147 psia1013 bar and 60 relative humidity which are established by the International Standards Organization ISO and frequently referred to as ISO conditions Air entering the compressor at point 1 is com pressed to some higher pressure No heat is added however compression raises the air temperature so that the air at the discharge of the compressor is at a higher temperature and pressure Upon leaving the compressor air enters the combustion system at point 2 where fuel is injected and combustion occurs The combus tion process occurs at essentially constant pres sure Although high local temperatures are reached within the primary combustion zone approaching stoichiometric conditions the GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 2 Mechanical Drive Gas Turbine Ratings Model Year ISO Rating ISO Rating Heat Heat Mass Mass Exhaust Exhaust Continuous Continuous Rate Rate Flow Flow Temp Temp kW hp Btushphr kJkWh lbsec kgsec degrees F degrees C M3142 J 1952 11290 15140 9500 13440 117 53 1008 542 M3142R J 1952 10830 14520 7390 10450 117 53 698 370 M5261 RA 1958 19690 26400 9380 13270 205 92 988 531 M5322R B 1972 23870 32000 7070 10000 253 114 666 352 M5352 B 1972 26110 35000 8830 12490 273 123 915 491 M5352R C 1987 26550 35600 6990 9890 267 121 693 367 M5382 C 1987 28340 38000 8700 12310 278 126 960 515 M6581 B 1978 38290 51340 7820 11060 295 134 1013 545 Table 2 GE gas turbine performance characteristics Mechanical drive gas turbine ratings MS7000 EA 12 PG Model Number of Shafts Power Series Application Approx Output Power in Hundreds Thousands or 10 Thousands of Horsepower R Regen Blank SC 1 or 2 Frame 357 69 Mech Drive Pkgd Gen M PG 1 7 Figure 1 Heavyduty gas turbine model designation GT25385A GT23054A combustion system is designed to provide mix ing burning dilution and cooling Thus by the time the combustion mixture leaves the com bustion system and enters the turbine at point 3 it is at a mixed average temperature In the turbine section of the gas turbine the energy of the hot gases is converted into work This conversion actually takes place in two steps In the nozzle section of the turbine the hot gases are expanded and a portion of the thermal energy is converted into kinetic energy In the subsequent bucket section of the turbine a portion of the kinetic energy is transferred to the rotating buckets and converted to work Some of the work developed by the turbine is used to drive the compressor and the remain der is available for useful work at the output flange of the gas turbine Typically more than 50 of the work developed by the turbine sec tions is used to power the axial flow compressor As shown in Figure 2 singleshaft gas turbines are configured in one continuous shaft and therefore all stages operate at the same speed These units are typically used for generator drive applications where significant speed varia tion is not required A schematic diagram for a simplecycle two shaft gas turbine is shown in Figure 3 The low pressure or power turbine rotor is mechani cally separate from the highpressure turbine and compressor rotor The low pressure rotor is said to be aerodynamically coupled This unique feature allows the power turbine to be operated at a range of speeds and makes two shaft gas turbines ideally suited for variable speed applications All of the work developed by the power turbine is available to drive the load equipment since the work developed by the highpressure tur bine supplies all the necessary energy to drive the compressor On twoshaft machines the starting requirements for the gas turbine load train are reduced because the load equipment is mechanically separate from the highpressure turbine The Brayton Cycle The thermodynamic cycle upon which all gas turbines operate is called the Brayton cycle Figure 4 shows the classical pressurevolume PV and temperatureentropy TS diagrams for this cycle The numbers on this diagram cor GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 3 Compressor Inlet Air 1 Combustor Fuel 2 4 Exhaust 3 Turbine Generator Figure 2 Simplecycle singleshaft gas turbine GT08922A respond to the numbers also used in Figure 2 Path 1 to 2 represents the compression occur ring in the compressor path 2 to 3 represents the constantpressure addition of heat in the combustion systems and path 3 to 4 represents the expansion occurring in the turbine The path from 4 back to 1 on the Brayton cycle diagrams indicates a constantpressure cooling process In the gas turbine this cooling is done by the atmosphere which provides fresh cool air at point 1 on a continuous basis in exchange for the hot gases exhausted to the atmosphere at point 4 The actual cycle is an open rather than closed cycle as indicated Every Brayton cycle can be characterized by two significant parameters pressure ratio and firing temperature The pressure ratio of the cycle is the pressure at point 2 compressor discharge pressure divided by the pressure at point 1 compressor inlet pressure In an ideal cycle GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 4 Exhaust Load LP Compressor Inlet Air Combustor Fuel HP Turbine Figure 3 Simplecycle twoshaft gas turbine 1 2 1 2 1 2 3 4 4 4 3 3 Fuel P T S V Figure 4 Brayton cycle GT08923C GT23055A this pressure ratio is also equal to the pressure at point 3 divided by the pressure at point 4 However in an actual cycle there is some slight pressure loss in the combustion system and hence the pressure at point 3 is slightly less than at point 2 The other significant parameter firing temper ature is thought to be the highest temperature reached in the cycle GE defines firing temper ature as the massflow mean total temperature at the stage 1 nozzle trailing edge plane Currently all first stage nozzles are cooled to keep the temperatures within the operating lim its of the materials being used The two types of cooling currently employed by GE are air and steam Air cooling has been used for more than 30 years and has been extensively developed in air craft engine technology as well as the latest fam ily of large power generation machines Air used for cooling the first stage nozzle enters the hot gas stream after cooling down the nozzle and reduces the total temperature immediately downstream GE uses this temperature since it is more indicative of the cycle temperature repre sented as firing temperature by point 3 in Figure 4 Steamcooled first stage nozzles do not reduce the temperature of the gas directly through mixing because the steam is in a closed loop As shown in Figure 5 the firing temperature on a turbine with steamcooled nozzles GEs cur rent H design has an increase of 200 degrees without increasing the combustion exit temperature An alternate method of determining firing tem perature is defined in ISO document 2314 Gas Turbines Acceptance Tests The firing tem perature here is a reference turbine inlet tem perature and is not generally a temperature that exists in a gas turbine cycle it is calculated from a heat balance on the combustion system using parameters obtained in a field test This ISO reference temperature will always be less than the true firing temperature as defined by GE in many cases by 100 F38 C or more for machines using air extracted from the compressor for internal cooling which bypasses the combustor Figure 6 shows how these various temperatures are defined GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 5 Figure 5 Comparison of aircooled vs steamcooled first stage nozzle OPEN LOOP AIRCOOLED NOZZLE ADVANCED CLOSED LOOP STEAMCOOLED NOZZLE 200F More Firing Temp at Same NOx Production Possible GT25134 Thermodynamic Analysis Classical thermodynamics permit evaluation of the Brayton cycle using such parameters as pres sure temperature specific heat efficiency fac tors and the adiabatic compression exponent If such an analysis is applied to the Brayton cycle the results can be displayed as a plot of cycle efficiency vs specific output of the cycle Figure 7 shows such a plot of output and efficiency for different firing temperatures and various pressure ratios Output per pound of airflow is important since the higher this value the smaller the gas turbine required for the same output power Thermal efficiency is important because it directly affects the operating fuel costs Figure 7 illustrates a number of significant points In simplecycle applications the top curve pressure ratio increases translate into efficiency gains at a given firing temperature GE Power Systems I GER3567H I 1000 6 GE Gas Turbine Performance Characteristics Turbine Inlet Temperature Average Gas Temp in Plane A TA ISO Firing Temperature Calculated Temp in Plane C TC fMa Mf GE Uses Firing Temperature TB Highest Temperature at Which Work Is Extracted Firing Temperature Average Gas Temp in Plane B TB CL Figure 6 Definition of firing temperature Figure 7 Gas turbine thermodynamics GT23056 GT17983A The pressure ratio resulting in maximum out put and maximum efficiency change with firing temperature and the higher the pressure ratio the greater the benefits from increased firing temperature Increases in firing temperature provide power increases at a given pressure ratio although there is a sacrifice of efficiency due to the increase in cooling air losses required to maintain parts lives In combinedcycle applications as shown in the bottom graph in Figure 7 pressure ratio increases have a less pronounced effect on effi ciency Note also that as pressure ratio increas es specific power decreases Increases in firing temperature result in increased thermal effi ciency The significant differences in the slope of the two curves indicate that the optimum cycle parameters are not the same for simple and combined cycles Simplecycle efficiency is achieved with high pressure ratios Combinedcycle efficiency is obtained with more modest pressure ratios and greater firing temperatures For example the MS7001FA design parameters are 2420 F1316 C firing temperature and 1571 pressure ratio while simplecycle efficiency is not maximized combinedcycle efficiency is at its peak Combined cycle is the expected application for the MS7001FA Combined Cycle A typical simplecycle gas turbine will convert 30 to 40 of the fuel input into shaft output All but 1 to 2 of the remainder is in the form of exhaust heat The combined cycle is generally defined as one or more gas turbines with heatrecovery steam generators in the exhaust producing steam for a steam turbine generator heattoprocess or a combination thereof Figure 8 shows a combined cycle in its simplest form High utilization of the fuel input to the gas turbine can be achieved with some of the more complex heatrecovery cycles involving multiplepressure boilers extraction or topping steam turbines and avoidance of steam flow to a condenser to preserve the latent heat content Attaining more than 80 utilization of the fuel input by a combination of electrical power gen eration and process heat is not unusual GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 7 Exhaust Fuel Gas Turbine Air Comp Turb Gen Gen Comb HRSG ST Turb Figure 8 Combined cycle GT05363C Gen Combined cycles producing only electrical power are in the 50 to 60 thermal efficien cy range using the more advanced gas turbines Papers dealing with combinedcycle applica tions in the GE Reference Library include GER3574F GE CombinedCycle Product Line and Performance GER3767 SingleShaft CombinedCycle Power Generation Systems and GER3430F Cogeneration Application Considerations Factors Affecting Gas Turbine Performance Air Temperature and Site Elevation Since the gas turbine is an airbreathing engine its performance is changed by anything that affects the density andor mass flow of the air intake to the compressor Ambient weather conditions are the most obvious changes from the reference conditions of 59 F15 C and 147 psia1013 bar Figure 9 shows how ambient tem perature affects the output heat rate heat con sumption and exhaust flow of a singleshaft MS7001 Each turbine model has its own tem peratureeffect curve as it depends on the cycle parameters and component efficiencies as well as air mass flow Correction for altitude or barometric pressure is more straightforward The air density reduces as the site elevation increases While the result ing airflow and output decrease proportionate ly the heat rate and other cycle parameters are not affected A standard altitude correction curve is presented in Figure 10 Humidity Similarly humid air which is less dense than dry air also affects output and heat rate as shown in Figure 11 In the past this effect was thought to be too small to be considered However with the increasing size of gas turbines and the utilization of humidity to bias water and steam injection for NOx control this effect has greater significance It should be noted that this humidity effect is a result of the control system approximation of firing temperature used on GE heavyduty gas turbines Singleshaft turbines that use turbine exhaust temperature biased by the compressor pressure ratio to the approximate firing tem perature will reduce power as a result of GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 8 7 0 130 Output Compressor Inlet Temperature Percent Design Heat Rate Exhaust Flow Heat Cons 120 49 100 38 80 27 60 16 40 4 20 18 F C 70 80 90 100 110 120 Figure 9 Effect of ambient temperature GT22045D increased ambient humidity This occurs because the density loss to the air from humidi ty is less than the density loss due to tempera ture The control system is set to follow the inlet air temperature function By contrast the control system on aeroderiva tives uses unbiased gas generator discharge tem perature to approximate firing temperature The gas generator can operate at different speeds from the power turbine and the power will actually increase as fuel is added to raise the moist air due to humidity to the allowable temperature This fuel increase will increase the gas generator speed and compensate for the loss in air density Inlet and Exhaust Losses Inserting air filtration silencing evaporative coolers or chillers into the inlet or heat recov ery devices in the exhaust causes pressure losses in the system The effects of these pressure loss es are unique to each design Figure 12 shows GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 9 Figure 10 Altitude correction curve Figure 11 Humidity effect curve GT18848B GT22046B the effects on the MS7001EA which are typical for the E technology family of scaled machines MS6001B 7001EA 9001E Fuels Work from a gas turbine can be defined as the product of mass flow heat energy in the com busted gas Cp and temperature differential across the turbine The mass flow in this equation is the sum of compressor airflow and fuel flow The heat energy is a function of the elements in the fuel and the products of combustion Tables 1 and 2 show that natural gas methane produces nearly 2 more output than does dis tillate oil This is due to the higher specific heat in the combustion products of natural gas resulting from the higher water vapor content produced by the higher hydrogencarbon ratio of methane This effect is noted even though the mass flow lbh of methane is lower than the mass flow of distillate fuel Here the effects of specific heat were greater than and in oppo sition to the effects of mass flow Figure 13 shows the total effect of various fuels on turbine output This curve uses methane as the base fuel Although there is no clear relationship between fuel lower heating value LHV and output it is possible to make some general assumptions If the fuel consists only of hydrocarbons with no inert gases and no oxygen atoms output increases as LHV increases Here the effects of Cp are greater than the effects of mass flow Also as the amount of inert gases is increased the decrease in LHV will provide an increase in output This is the major impact of IGCC type fuels that have large amounts of inert gas in the fuel This mass flow addition which is not com pressed by the gas turbines compressor increases the turbine output Compressor power is essentially unchanged Several side effects must be considered when burning this kind of lower heating value fuels I Increased turbine mass flow drives up compressor pressure ratio which eventually encroaches on the compressor surge limit I The higher turbine power may exceed fault torque limits In many cases a larger generator and other accessory equipment may be needed I High fuel volumes increase fuel piping and valve sizes and costs Low or mediumBtu coal gases are frequently supplied at high temperatures which further increases their volume flow GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 10 Figure 12 Pressure drop effects MS7001EA 4 Inches 10 mbar H2O Inlet Drop Produces 142 Power Output Loss 045 Heat Rate Increase 19 F 11 C Exhaust Temperature Increase 4 Inches 10 mbar H2O Exhaust Drop Produces 042 Power Output Loss 042 Heat Rate Increase 19 F 11 C Exhaust Temperature Increase GT18238C I LowerBtu gases are frequently saturated with water prior to delivery to the turbine This increases the combustion products heat transfer coefficients and raises the metal temperatures in the turbine section which may require lower operating firing temperature to preserve parts lives I As the Btu value drops more air is required to burn the fuel Machines with high firing temperatures may not be able to burn low Btu gases I Most airblown gasifiers use air supplied from the gas turbine compressor discharge I The ability to extract air must be evaluated and factored into the overall heat and material balances As a result of these influences each turbine model will have some application guidelines on flows temperatures and shaft output to preserve its design life In most cases of operation with lower heating value fuels it can be assumed that output and efficiency will be equal to or higher than that obtained on natural gas In the case of higher heating value fuels such as refinery gases output and efficiency may be equal to or lower than that obtained on natural gas Fuel Heating Most of the combined cycle turbine installations are designed for maximum efficiency These plants often utilize integrated fuel gas heaters Heated fuel results in higher turbine efficiency due to the reduced fuel flow required to raise the total gas temperature to firing temperature Fuel heating will result in slightly lower gas tur bine output because of the incremental volume flow decrease The source of heat for the fuel typically is the IP feedwater Since use of this energy in the gas turbine fuel heating system is thermodynamically advantageous the com bined cycle efficiency is improved by approxi mately 06 GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 11 100 H2 40 20 100 CH4 Output Percent Kcalkg Thousands LHVBtulb Thousands 105 100 CO 75 N2 25 CH4 75 CO2 25 CH4 100 CH4H10 60 50 30 20 10 10 30 0 100 110 115 120 125 130 Figure 13 Effect of fuel heating value on output GT25842 Diluent Injection Since the early 1970s GE has used water or steam injection for NOx control to meet appli cable state and federal regulations This is accomplished by admitting water or steam in the cap area or headend of the combustion liner Each machine and combustor configura tion has limits on water or steam injection levels to protect the combustion system and turbine section Depending on the amount of water or steam injection needed to achieve the desired NOx level output will increase because of the additional mass flow Figure 14 shows the effect of steam injection on output and heat rate for an MS7001EA These curves assume that steam is free to the gas turbine cycle therefore heat rate improves Since it takes more fuel to raise water to combustor conditions than steam water injection does not provide an improve ment in heat rate Air Extraction In some gas turbine applications it may be desirable to extract air from the compressor Generally up to 5 of the compressor airflow can be extracted from the compressor dis charge casing without modification to casings or onbase piping Pressure and air temperature will depend on the type of machine and site conditions Air extraction between 6 and 20 may be possible depending on the machine and combustor configuration with some modi fications to the casings piping and controls Such applications need to be reviewed on a casebycase basis Air extractions above 20 will require extensive modification to the tur bine casing and unit configuration Figure 15 shows the effect of air extraction on output and heat rate As a rule of thumb every 1 in air extraction results in a 2 loss in power Performance Enhancements Generally controlling some of the factors that affect gas turbine performance is not possible The planned site location and the plant config uration such as simple or combinedcycle determine most of these factors In the event additional output is needed several possibilities to enhance performance may be considered GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 12 130 120 110 100 90 80 70 Output Compressor Inlet Temperature 40 60 80 100 120 ºF 4 16 27 38 49 ºC 1 3 No Steam Injection With 5 Steam Injection Figure 14 Effect of steam injection on output and heat rate Figure 15 Effect of air extraction on output and heat rate GT18851A GT220481C Inlet Cooling The ambient effect curve see Figure 9 clearly shows that turbine output and heat rate are improved as compressor inlet temperature decreases Lowering the compressor inlet tem perature can be accomplished by installing an evaporative cooler or inlet chiller in the inlet ducting downstream of the inlet filters Careful application of these systems is necessary as con densation or carryover of water can exacerbate compressor fouling and degrade performance These systems generally are followed by mois ture separators or coalescing pads to reduce the possibility of moisture carryover As Figure 16 shows the biggest gains from evap orative cooling are realized in hot lowhumid ity climates It should be noted that evapora tive cooling is limited to ambient temperatures of 59 F15 C and above compressor inlet tem perature 45 F72 C because of the potential for icing the compressor Information con tained in Figure 16 is based on an 85 effective evaporative cooler Effectiveness is a measure of how close the cooler exit temperature approaches the ambient wet bulb tempera ture For most applications coolers having an effectiveness of 85 or 90 provide the most economic benefit Chillers unlike evaporative coolers are not lim ited by the ambient wet bulb temperature The achievable temperature is limited only by the capacity of the chilling device to produce coolant and the ability of the coils to transfer heat Cooling initially follows a line of constant GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 13 Figure 16 Effect of evaporative cooling on output and heat rate 49 120 F C Dry Bulb Temperature 20 25 30 35 40 100 RH 15 005 020 Specific Humidity Evaporative Cooling Process Btu Per Pound of Dry Air Simplified Psychrometric Chart 000 010 015 40 60 80 100 4 16 27 38 10 RH Inlet Chilling Process 20 RH 40 RH 60 RH Figure 17 Inlet chilling process GT224191D GT21141D specific humidity as shown in Figure 17 As satu ration is approached water begins to condense from the air and mist eliminators are used Further heat transfer cools the condensate and air and causes more condensation Because of the relatively high heat of vaporization of water most of the cooling energy in this regime goes to condensation and little to temperature reduction Steam and Water Injection for Power Augmentation Injecting steam or water into the head end of the combustor for NOx abatement increases mass flow and therefore output Generally the amount of water is limited to the amount required to meet the NOx requirement in order to minimize operating cost and impact on inspection intervals Steam injection for power augmentation has been an available option on GE gas turbines for over 30 years When steam is injected for power augmentation it can be introduced into the compressor discharge casing of the gas turbine as well as the combustor The effect on output and heat rate is the same as that shown in Figure 14 GE gas turbines are designed to allow up to 5 of the compressor airflow for steam injec tion to the combustor and compressor dis charge Steam must contain 50 F28 C super heat and be at pressures comparable to fuel gas pressures When either steam or water is used for power augmentation the control system is normally designed to allow only the amount needed for NOx abatement until the machine reaches base full load At that point additional steam or water can be admitted via the governor control Peak Rating The performance values listed in Table 1 are base load ratings ANSI B1336 Ratings and Performance defines base load as operation at 8000 hours per year with 800 hours per start It also defines peak load as operation at 1250 hours per year with five hours per start In recognition of shorter operating hours it is possible to increase firing temperature to gen erate more output The penalty for this type of operation is shorter inspection intervals Despite this running an MS5001 MS6001 or MS7001 at peak may be a costeffective way to obtain more kilowatts without the need for additional peripheral equipment Generators used with gas turbines likewise have peak ratings that are obtained by operating at higher power factors or temperature rises Peak cycle ratings are ratings that are customized to the mission of the turbine considering both starts and hours of operation Firing tempera tures between base and peak can be selected to maximize the power capabilities of the turbine while staying within the starts limit envelope of the turbine hot section repair interval For instance the 7EA can operate for 24000 hours on gas fuel at base load as defined The starts limit to hot section repair interval is 800 starts For peaking cycle of five hours per start the hot section repair interval would occur at 4000 hours which corresponds to operation at peak firing temperatures Turbine missions between five hours per start and 800 hours per start may allow firing temperatures to increase above base but below peak without sacrificing hours to hot section repair Water injection for power aug mentation may be factored into the peak cycle rating to further maximize output Performance Degradation All turbomachinery experiences losses in per formance with time Gas turbine performance degradation can be classified as recoverable or nonrecoverable loss Recoverable loss is usually GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 14 associated with compressor fouling and can be partially rectified by water washing or more thoroughly by mechanically cleaning the com pressor blades and vanes after opening the unit Nonrecoverable loss is due primarily to increased turbine and compressor clearances and changes in surface finish and airfoil con tour Because this loss is caused by reduction in component efficiencies it cannot be recovered by operational procedures external mainte nance or compressor cleaning but only through replacement of affected parts at rec ommended inspection intervals Quantifying performance degradation is diffi cult because consistent valid field data is hard to obtain Correlation between various sites is impacted by variables such as mode of opera tion contaminants in the air humidity fuel and dilutent injection levels for NOx Another prob lem is that test instruments and procedures vary widely often with large tolerances Typically performance degradation during the first 24000 hours of operation the normally recommended interval for a hot gas path inspection is 2 to 6 from the performance test measurements when corrected to guaran teed conditions This assumes degraded parts are not replaced If replaced the expected per formance degradation is 1 to 15 Recent field experience indicates that frequent offline water washing is not only effective in reducing recoverable loss but also reduces the rate of nonrecoverable loss One generalization that can be made from the data is that machines located in dry hot cli mates typically degrade less than those in humid climates Verifying Gas Turbine Performance Once the gas turbine is installed a perform ance test is usually conducted to determine power plant performance Power fuel heat consumption and sufficient supporting data should be recorded to enable astested per formance to be corrected to the condition of the guarantee Preferably this test should be done as soon as practical with the unit in new and clean condition In general a machine is considered to be in new and clean condition if it has less than 200 fired hours of operation Testing procedures and calculation methods are patterned after those described in the ASME Performance Test Code PTC221997 Gas Turbine Power Plants Prior to testing all sta tion instruments used for primary data collec tion must be inspected and calibrated The test should consist of sufficient test points to ensure validity of the test setup Each test point should consist of a minimum of four complete sets of readings taken over a 30minute time period when operating at base load Per ASME PTC22 1997 the methodology of correcting test results to guarantee conditions and measurement uncertainties approximately 1 on output and heat rate when testing on gas fuel shall be agreed upon by the parties prior to the test Summary This paper reviewed the thermodynamic princi ples of both one and twoshaft gas turbines and discussed cycle characteristics of the several models of gas turbines offered by GE Ratings of the product line were presented and factors affecting performance were discussed along with methods to enhance gas turbine output GE heavyduty gas turbines serving industrial utility and cogeneration users have a proven history of sustained performance and reliabili ty GE is committed to providing its customers with the latest in equipment designs and advancements to meet power needs at high thermal efficiency GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 15 List of Figures Figure 1 Heavyduty gas turbine model designation Figure 2 Simplecycle singleshaft gas turbine Figure 3 Simplecycle twoshaft gas turbine Figure 4 Brayton cycle Figure 5 Comparison of aircooled vs steamcooled first stage nozzle Figure 6 Definition of firing temperature Figure 7 Gas turbine thermodynamics Figure 8 Combined cycle Figure 9 Effect of ambient temperature Figure 10 Altitude correction curve Figure 11 Humidity effect curve Figure 12 Pressure drop effects MS7001EA Figure 13 Effect of fuel heating value on output Figure 14 Effect of steam injection on output and heat rate Figure 15 Effect of air extraction on output and heat rate Figure 16 Effect of evaporative cooling on output and heat rate Figure 17 Inlet chilling process List of Tables Table 1 GE gas turbine performance characteristics Generator drive gas turbine ratings Table 2 GE gas turbine performance characteristics Mechanical drive gas turbine ratings GE Gas Turbine Performance Characteristics GE Power Systems I GER3567H I 1000 16