The combined cycle and variations that use HRSGs
Joseph Miller, The Energy Corporation, Steamboat Springs, CO, United States
Abstract
The Brayton cycle (gas turbine) and the Rankine cycle (steam turbine) are two venerable cycles that have served mankind well. However, the combined cycle, which combines the Brayton and Rankine cycles, has resulted in cycle efficiencies exceeding 60% on a lower heating value basis. This is a much higher efficiency than can be achieved by either the Brayton or Rankine cycle alone. To combine these two cycles, a means to recover the waste heat from the gas turbine exhaust must be provided. The modern day heat recovery steam generator (HRSG) has met this need and is the bridge between the two cycles. The versatility of the modern day HRSG has allowed great flexibility in combined cycle design: single pressure, double pressure, triple pressure steam levels; nonreheat, reheat; and supplementary firing. In addition, the adaptability in HRSG design has provided the prerequisite heat recovery for variants of the combined cycle such as cogeneration, steam power augmentation, integrated gasification combined cycle, and solar hybrid cycles. Without the HRSG, the combined cycle and its variants would not be technically feasible.
Keywords
Brayton cycle; Rankine cycle; combined cycle; heat recovery steam generators; HRSGs; HRSGs in combined cycle design; pressure levels; gas turbines
2.1 Introduction
Without question, energy—or more precisely, the consumption of energy—drives the world economy. We search the depths of the sea for oil to refine into various grades of fuel to power aircraft engines, trucks, and automobiles. We mine for coal on all corners of the globe to combust this fuel source to generate electricity and produce steel. We split atoms of radioactive substances, unleashing enormous amounts of nuclear energy from a relatively small amount of mass. We fracture underground shale deposits to harvest natural gas for use as an industrial feedstock, to heat homes and water, and to generate electricity. We harness the wind, we use the sun’s radiation—we even try to capture the force of ocean tides—to meet mankind’s collective, unyielding demand for energy. But this needs qualification. The world economy demands not just energy, but inexpensive energy, especially inexpensive electricity.
It is with this global, unchanging backdrop that we explore combined cycle power plants, and other power cycle variants, that use heat recovery steam generators (HRSGs). HRSGs fill a unique role in the neverending quest for inexpensive electricity to power the world.
2.2 Combining the Brayton and Rankine cycles
The Brayton cycle is synonymous with the modern day gas turbine but that is not how it started. Named after American engineer George Brayton (1830–92), the cycle was first proposed by Englishman John Barber in the late 1700s. As developed by Brayton, the machine was a constant pressure reciprocating engine constructed of separate piston compressor and expander sections. Compressed air was heated by combusting a vaporized fuel; useful work, such as driving a water pump or textile mill, was performed during the expansion process.
Fig. 2.1 depicts the ideal or fully reversible (no entropy production) Brayton cycle plotted on a temperature–entropy diagram. Comprised of two adiabatic-reversible and two constant pressure processes, this cycle has evolved into an integral component of the world economy. The modern day Brayton cycle efficiently and reliably powers airplanes and ships, and is used to generate electricity. In its ideal cycle form, gas is isentropically compressed from Point 1 to Point 2, followed by a constant pressure heat addition (Point 2 to Point 3) raising the working gas temperature. The gas then isentropically expands from Point 3 to Point 4. To close the ideal cycle, the working gas undergoes a constant pressure cooling process (Point 4 to Point 1), returning to Point 1 to restart the cycle at the original state point.
In its modern form (i.e., the gas turbine), the Brayton cycle is built from three major components: a multistage, axial compressor; one or more combustion chambers (called combustors); and a turbine for expanding the working gas. Fig. 2.2 below illustrates these three components of an open cycle gas turbine driving a generator for electricity production. It is an open cycle because unlike the ideal Brayton cycle shown in Fig. 2.1, the working gas is not cooled; rather, it is discharged to the atmosphere after expanding through the turbine. Comparing Figs. 2.1 and 2.2, note the compression of air from Point 1 to Point 2, the heating of the compressed air by the addition of a vaporized fuel in the combustor from Point 2 to Point 3, then the expansion of the high temperature and high pressure air/fuel mixture through the turbine from Point 3 to Point 4. The air/fuel mixture, as previously mentioned, does not return to state point 1.
What has just been described and depicted in Fig. 2.2 is a simple cycle gas turbine generator used predominately for peaking power service. The Brayton cycle turbine spins the generator to produce electricity. Depending on the generator rotational speed measured in revolutions per minute (rpm), either 50 or 60 Hz electricity is produced. Simple cycle “peakers,” as they are known in the electrical power industry, can reach full power output in less than 10 minutes. This is a critically important capability during electrical grid disturbances where additional power generation is required to prevent grid underfrequency and possible blackout events. But the exhaust gas, after expanding through the turbine, is discharged to the atmosphere at temperatures typically in excess of 1000°F. As we will discuss shortly, this wastes significant amounts of energy that could still be captured to produce useful work.
Around the mid-1800s, a Scottish civil engineering professor named William J.M. Rankine is credited with describing an ideal vapor–liquid cycle that is unquestionably recognized as the precursor to the modern day steam power plant. In Rankine’s ideal cycle, shown diagrammatically on the temperature–entropy diagram in Fig. 2.3, the vapor and liquid undergo a phase change by the addition and subtraction of heat.
At Point 1 the working fluid isentropically expands to a lower pressure at Point 2 and in the process reduces in temperature while performing work. The working fluid undergoes a constant pressure cooling process from Point 2 to Point 3. A phase change from a saturated two-phase substance to a fully liquid state occurs in the cooling process. Point 3 to Point 4 consists of an isentropic compression of the working fluid followed by a constant pressure heat addition from Point 4 to Point 1. This ideal closed cycle represents any working fluid that undergoes a phase change. Brilliantly for mankind, the Rankine cycle has been developed using water as the working fluid to generate electricity since the late 1880s, only thirty-some years from the time William Rankine described his heat engine cycle.
In Rankine cycle power plants, superheated steam is expanded through a steam turbine driving an electrical generator (Point 1 to Point 2). Heat is rejected in a condenser that turns the two-phase mixture back to water (Point 2 to Point 3). Pumps are used to feed water into the steam boiler at the desired pressure (Point 3 to Point 4). Fuel is combusted in the boiler to supply the heat required to change the water back to superheated steam.
The fuel flexibility of the steam Rankine cycle is tremendous. Boilers have been, and continue to be, fired on coal, oil, natural gas, wood, other biomass, refuse-derived fuel, even shredded tires. Nuclear power plants are based on the Rankine cycle, with the splitting of atoms providing the heat source. Tapping into the heat of the earth’s inner core, geothermal power plants use vapor or liquid-dominated resources to spin steam turbines for electrical generation. Organic Rankine cycles use a low boiling point, carbon-based, working fluid to capture low-grade heat and convert it into electricity. The Rankine cycle is even adaptable to use the sun’s radiation to heat a working fluid and generate electricity in concentrated solar power plants (CSP).
We have described two fundamentally very different cycles to generate electricity: the Brayton cycle, which predominately uses an air/fuel mixture as the working fluid, and the Rankine cycle, which predominately uses water as the vapor–liquid working fluid. Air and water are two very abundant earth resources. The crux of the problem is fossil fuel, being finite, is subject to the forces of supply and demand pricing. Generating electricity inexpensively then must be done efficiently. So what would happen if we combined the two power cycles? How much more efficient could this combined cycle be compared to the Brayton and Rankine cycles separately? And how do we combine the cycles? What piece of equipment would be necessary?
Remember that the turbine exhaust gas from a simple cycle gas turbine discharges to the atmosphere. This exhaust stream is still at a high temperature albeit at a low pressure. The waste heat available in the turbine exhaust gas can be recovered. Early concepts considered using the gas turbine exhaust in combination with additional combustion air to burn a fuel source in a boiler. This would generate steam for use in a Rankine cycle. But advancements in gas turbine firing temperature (Point 3 of the Brayton cycle) soon yielded turbine exhaust gas temperatures (Point 4) hot enough to directly generate steam at suitable temperatures for the steam turbine. The gas turbine (i.e., Brayton cycle) then becomes the “topping cycle” and the steam turbine (i.e., Rankine cycle) becomes the “bottoming cycle.” With this arrangement, the modern combined cycle was born, with the HRSG providing the means to capture the waste heat from the gas turbine.
Fig. 2.4 provides a schematic of a combined cycle power plant. State points have been modified with a “B” for the Brayton cycle and “R” for the Rankine cycle. The turbine exhaust gas at Point B4 enters into the HRSG to heat feedwater and produce steam, with the exhaust gas then exiting the stack at Point B4’ at a significantly reduced temperature. A single pressure level HRSG is shown simply for clarity. As will be seen later in this chapter, HRSGs are intricately more complex than the representation depicted in Fig. 2.4.
Combining the Brayton and Rankine cycles created the need for a new piece of power plant equipment: the HRSG. Today’s HRSG is the bridge between the two fundamentally different power cycles. And like a physical bridge connecting two different towns allowing each town to benefit from the other, the HRSG connects the two distinct power cycles yielding a large improvement in thermal efficiency compared to each cycle by itself.
2.3 The central role of HRSGs in combined cycle design
The world’s first gas turbine for electrical power generation reportedly began operation in Europe in 1939. Ten years later, the first combined cycle power plant in the United States entered into service in Oklahoma City. Oklahoma Gas & Electric’s Belle Isle Station had, by today’s standards, a small 3.5 MW gas turbine generator and used the turbine exhaust to heat boiler feedwater.
Modern combined cycle power plants have gas turbines ranging in size from single-digit megawatts to in excess of 500 MW. Turbine exhaust gas temperatures and exhaust flow rates have continually increased as gas turbine manufacturers strive for higher efficiencies and greater power density.
Central to the success of combined cycle power plants has been the ability of HRSG design to evolve in step with the gas turbine. As gas turbines became larger, HRSGs became larger to handle the increase in exhaust gas flow. As gas turbine firing temperature increased, HRSG heat transfer metallurgy and design adapted to successfully contend with the increase in turbine exhaust gas temperatures. As natural gas prices increased and even higher efficiencies were required to lower the cost of electricity production, reheat capability was introduced into HRSG design. Because gas turbine power output and exhaust flow decreases at hotter ambient dry bulb temperatures, supplementary firing capability was added to HRSGs to provide capacity stabilization. Single pressure level HRSG design gave way to two-pressure nonreheat, which in turn gave way to three-pressure, reheat HRSGs with ever higher high-pressure (HP) and reheat steam temperatures. This adaptability has time and again proven the unique and central role HRSGs perform in combining the Brayton and Rankine cycles.
2.3.1 Pressure levels
Gas turbines for power generation applications can be categorized into two distinct groups: aeroderivative engines and industrial heavy frame machines. Aeroderivative gas turbines, as the name implies, were derived from aircraft jet engines. Lightweight and fast starting, aeroderivatives have power outputs up to 100 MW. The most efficient aeroderivatives in simple cycle applications are just over 40% on a lower heating value (LHV) fuel basis. Heavy frame gas turbines were developed specifically for mechanical drive and power generation service. These gas turbines have an extremely large power output range—from single-digit MW units to engines over 500 MW in 50 Hz service. The most efficient heavy frame machines are also over 40% LHV efficiency.
The need for the wide range in gas turbine power outputs is apparent. This output variability provides the ability to precisely match the load requirements. And the need for high thermal efficiency is also readily apparent: higher efficiency means less fuel burn per megawatt-hour of electrical energy production and lower electricity production costs. But how does this impact HRSG design, and more specifically, the number of pressure levels in the HRSG? To answer this question, it is important to understand how the air/fuel mixture temperature at Point 3 of the Brayton cycle (i.e., the gas turbine firing temperature) impacts gas turbine efficiency.
The work done in the expansion turbine of the Brayton cycle is equal to the rate of change in the working fluid’s enthalpy. This can be expressed by the following equation:
where:
H is the total enthalpy of the working fluid, which is in part a function of temperature.
The above equation can be also expressed as:
where:
m is the mass rate and h is the specific enthalpy of the working fluid.
The net power output of a gas turbine (Wn) is equal to the turbine section work minus the power necessary for the compressor section. By numerous variable substitutions and equation rewrites, the gas turbine net power output can be expressed as:
cp is the specific heat at constant pressure and k is the ratio of specific heats,
T1 and T3 are the ambient and firing temperatures,
ƞT and ƞC are the polytropic efficiencies of the turbine and compressor sections respectively.
From the equation, the net power output of the gas turbine increases as the T3 firing temperature increases. Therefore, for a given amount of heat added to the cycle, as state Point 3 temperature increases, the gas turbine efficiency also increases. In the ideal world, gas turbine firing temperatures would approach stoichiometric combustion temperatures. The turbine inlet temperature in the real world is limited by metallurgy. At some point, the turbine blades would oxidize, yield, and fail due to excessive temperatures. Fortunately, gas turbine manufacturers have been able to design and manufacture turbine blades with air and steam cooling as well as coatings that have pushed the latest model turbine inlet temperatures to 2900°F. This is in excess of the melting point of carbon steel, stainless steels, and Inconel.
For a given compression ratio, an increase in state Point 3 temperature results in a corresponding increase of state Point 4 temperature. Hence, as gas turbine manufacturers have increased firing temperature over the years to improve efficiency, the turbine exhaust gas temperature has also increased (see Fig. 2.5 below). From the very first gas turbine in power plant application to the present heavy frame models, exhaust gas temperatures have increased nearly fourfold from roughly 550°F to 1200°F. Considering present state-of-the-art HP and reheat steam temperatures in the Rankine cycle are slightly higher than 1100°F, gas turbines make an ideal topping cycle for the combined cycle power plant.
The progression of gas turbine exhaust flow over the years has also been remarkable. Fig. 2.6 is a graph of the turbine exhaust flow for the largest heavy frame gas turbines commercially available in each time period for the 60 Hz market. From the late 1970s to the present, turbine exhaust flow has nearly doubled in a fairly linear progression.
High turbine exhaust flow rates at high temperatures yield a significant amount of energy for the bottoming cycle. The key to the HRSG’s ability to effectively capture the topping cycle waste heat as the exhaust energy has progressively increased has been through the addition of pressure levels within the HRSG. Fig. 2.7 provides a typical temperature profile of the turbine exhaust gas and the water-steam working fluids within the HRSG. A single pressure level comprised of an economizer, an evaporator section, and a superheater is depicted.
Feedwater enters the economizer and is heated by the exhaust gas. The water temperature increases and approaches the saturation temperature of the evaporator section pressure. After entering the evaporator section, the water boils, creating a steam/water mixture. The temperature of the steam/water mixture remains constant during the phase change. The heat to boil the water and generate steam is provided by the exhaust gas as it flows past the evaporator section tubes (the exhaust gas flows externally to the tubes, steam/water flows through the inside of the tubes). As the exhaust gas exits the evaporator section of the HRSG, its temperature must be higher than the saturation temperature of the steam/water mixture by what is known as the “pinch” temperature. Heat transfer can only occur if the heat source is at a higher temperature than the fluid being heated. For the exhaust gas temperature to equal the saturation temperature of the steam/water mixture an infinite amount of heat transfer surface area would be required. Typical pinch temperatures are 14°F to 20°F based on reasonable economic considerations. The last HRSG section shown is the superheater. Here the steam generated in the evaporator section is increased in temperature (i.e., is superheated).
A single pressure level in the HRSG cannot economically capture all of the available gas turbine waste heat for reasons that will be explained in detail in Chapter 3. Even if the pinch temperature is reduced to zero and a superheater section is part of the single-pressure HRSG design, not all of the available waste heat will be recovered. The HRSG stack temperature will still be relatively high.
One solution for increasing the energy recovery in the HRSG has been to add pressure levels. Instead of just one pressure level, the HRSG can generate steam at two or three different pressures. This has worked well since the steam turbines used in combined cycle power plants can readily accommodate either two or three steam pressure admissions. For nonreheat cycles, steam generated in the HRSG can be admitted in the steam turbine as shown in Fig. 2.8.
In a two-pressure nonreheat cycle, HP steam and low-pressure (LP) steam generated in the HRSG are admitted to the HP/IP and LP sections of the steam turbine respectively. For a three-pressure nonreheat cycle, IP steam is sent to the intermediate pressure (IP) steam turbine section in addition to the HP and LP steam flows previously shown in the two-pressure design.
Fig. 2.9 represents the standard three-pressure reheat cycle configuration for combined cycle power plants. Similar to the nonreheat steam turbine, HP steam and LP steam are directly admitted to the steam turbine. However, note that the exhaust steam from the HP section of the steam turbine is routed back to the HRSG for “reheating.” This steam flow is also referred to as cold reheat steam. Prior to entering into the reheater section of the HRSG, the cold reheat steam is combined with IP steam generated from the HRSG. This combined steam flow is heated in the HRSG reheater, then routed to the IP steam turbine admission port as hot reheat steam. The benefit of reheat will be discussed in Section 2.3.2.
Illustrated in Fig. 2.10 is a three-pressure HRSG showing only the evaporator section for each pressure level. Shown in Fig. 2.10 is the exhaust gas temperature leaving each evaporator section (HP=high pressure; IP=intermediate pressure; LP=low pressure) based on a 15°F pinch for each evaporator pressure. The saturation pressure used for each pressure level is representative of present day combined cycle power plants with large, heavy frame engines. Note the cascading exhaust gas temperature in the direction of exhaust gas flow. Clearly if only one pressure level is used, the exhaust gas temperature leaving the HRSG would be too high considering the importance of cycle efficiency in generating low-cost electricity.
To summarize, gas turbine manufacturers have continually raised engine firing temperature to improve gas turbine efficiency. Higher firing temperatures result in higher turbine exhaust gas temperatures. When coupled with the increase in turbine exhaust flow of the latest gas turbine models, a tremendous amount of waste heat is available for recovery in the HRSG. One means of capturing more of the waste heat, thereby improving overall combined cycle efficiency, is to add pressure levels to the HRSG. This HRSG design technique has been very effective, such that three pressure levels are the norm for combined cycle power plants. We will now turn our attention to another means of improving cycle efficiency within the HRSG.
2.3.2 Reheat
The Carnot cycle is an ideal cycle. It contains all fully reversible processes (see Fig. 2.11). In this cycle there are no friction losses; there is no destruction in availability, hence no entropy production. Each state point returns to exactly the same place from whence it started. The Carnot cycle, due to its fully reversible nature, represents the highest cycle efficiency possible for the two temperature limits of TH and TL; where TH represents both the heat source temperature and the temperature of the working fluid, and TL is both the working fluid temperature and the temperature of the heat sink.
In the real word, there are friction losses in pipe. Steam and water flow from high pressure to lower pressure and cannot reverse their path unless additional energy is consumed. There are unrecoverable losses when steam is throttled across a valve. Once a fuel is combusted it cannot return to its previous state. In the real world these processes are irreversible. Entropy is increased.
Heat transfer in the Carnot cycle occurs at zero temperature differential, an impossibility in the real world. For heat to transfer from one fluid to another, there must be a temperature difference, one fluid hotter than the other. During the heat transfer process no work is performed between the two fluids. One is simply increasing the temperature of the lower temperature fluid. Heat transfer is also irreversible. The hotter fluid giving up heat cannot return to its original temperature without additional energy being consumed. The larger the temperature difference, the larger the irreversibility. The larger the irreversibility, the larger the loss in availability—and the larger the reduction in efficiency. The goal then to improve cycle efficiency is to minimize the temperature difference between the heat source and working fluid. This holds true regardless of the heat source, be it combustion gases in a boiler or waste heat from a gas turbine exhaust stream.
Employing reheat is one means to reduce the temperature differential between the heat source and working fluid. Referring back to Fig. 2.9, HP steam, after expanding through the HP turbine section, is returned to the HRSG so the steam temperature can be increased (i.e., “reheated”). By reheating the steam, the composite temperature difference between the heat source (gas turbine exhaust) and the working fluid (steam/water) is reduced.
A single reheat cycle is shown in Fig. 2.12. Pressure losses (friction) are assumed to be zero (i.e., constant pressure heat addition). HP steam expands through the HP turbine section (Point 1 to Point 2) and then is returned to the HRSG for reheating (Point 2 to Point 3). The hot reheat steam is then expanded through the IP and LP steam turbine sections (Point 3 to Point 4). Point 4 to Point 5 is the constant pressure cooling process, Point 5 to Point 6 is the feedwater pumping process, and Point 6 to Point 1 is the initial heating step.
It stands to reason then that if one stage of reheat improves overall cycle efficiency then two or more stages of reheat would improve efficiency even more and be a sound economic choice. In theory yes, but in reality, no. Additional reheat stages soon experience diminishing returns. Unlike the ideal cycle where piping losses are ignored, routing steam back and forth between the HRSG and the steam turbine results in pressure loss, which is irreversible. Further, the additional steam piping, valves, instrumentation, and insulation for the reheat piping increases construction costs. The additional capital cost of more than one or two stages of reheat, in conjunction with the added complexity, has not been economically viable. To date, only single reheat has been employed for combined cycle power plants.
With respect to reheat pressure drops and implementing a single stage of reheat into a combined cycle power plant, it is important to keep the total pressure drop of the reheat piping and HRSG reheater modules to 10% or less of the HP turbine exhaust pressure. This design rule yields reasonable cold and hot reheat piping diameters while maximizing the gain in efficiency from employing reheat.
Another tangible benefit of reheat is its impact on steam quality in the last stages of the LP turbine. Since reheat increases the temperature of steam entering the IP steam turbine section, the steam moisture level is lower in the L-1 and L-0 (last two rows) turbine blades. This reduces blade moisture losses, which slightly improves cycle efficiency. The drier steam also reduces blade leading edge erosion.
2.3.3 Other decisions affecting heat recovery
HRSGs in combined cycle power plants are an amazing bridge between the Brayton and Rankine cycles. By adding pressure levels, maximum heat recovery can be achieved, while creating different steam pressures for smooth integration with the steam turbine. By employing a single reheat stage within the HRSG, the overall cycle efficiency can be increased by reducing irreversible cycle losses. But there are other HRSG design decisions that also affect heat recovery, and hence, cycle efficiency. Four of the major design decisions are briefly discussed below.
2.3.3.1 Amount of surface area
Without question, the amount of heat transfer surface area included in the HRSG has the biggest impact on the amount of heat recovered. Even if the HRSG has three pressure levels and one stage of reheat, without sufficient surface area, energy will be wasted up the stack and lost. Once the exhaust gas mixes with the atmosphere, the heat is unrecoverable.
The basic equation governing heat transfer in the HRSG is:
Q is the amount of heat transferred;
U is the overall heat transfer coefficient;
The amount of heat transferred, therefore, is directly a function of the total amount of heat transfer surface area included in the HRSG. With a multipressure HRSG, the amount of surface area for each pressure level must be determined. Since HP steam has the highest availability to do work, the amount of HP surface area is typically maximized within the previously discussed constraints of the evaporator pinch. Adding HP evaporator surface area to achieve a pinch of less than 14°F becomes very costly. Sufficient superheater and reheater surface area must be selected to achieve the desired steam temperatures. Too much economizer surface area can lead to steaming economizer problems.
2.3.3.2 Surface area sequencing
Surface area sequencing refers to how the different sections within a pressure level (economizer, evaporator, superheater) are arranged between the different pressure levels. Clearly, for each pressure level, feedwater must first be heated in the economizer section to raise the subcooled liquid’s temperature close to saturation temperature, then sent to the evaporator tubes to boil the feedwater and generate steam. From the evaporator, the saturated steam enters the superheater to raise the steam to the desired steam temperature. To obtain the desired steam temperatures for the hottest steam (HP steam and hot reheat steam), the HP superheater and reheater sections must be in the front of the HRSG (front being defined as the end closest to the gas turbine exhaust flange). This is where the exhaust gas temperature is highest. Typically, the HP superheater and reheater are split into at least two different sections each. This allows locating an attemperator between the split sections for temperature control. Depending on the desired IP steam and LP steam temperatures, more than one superheater for each of these pressure levels may be required with the finishing superheater colocated with a higher pressure surface area section where the exhaust gas temperature is hotter.
2.3.3.3 Supplementary firing
The gas turbine exhaust gas has a sufficient oxygen concentration to support supplementary firing within the HRSG. Supplementary firing or “duct firing” consists of injecting an additional fuel source inside the HRSG to mix with the turbine exhaust gas stream, where it is then ignited to increase the energy content of the exhaust gas. Duct firing can double the HP steam production at base load of the gas turbine. The practical limit for duct firing is around 1600°F to 1650°F bulk gas temperature measured downstream of the combustion zone but upstream of the first downstream surface area from the duct burner.
Figs. 2.13 and 2.14 show two potential duct burner locations within the HRSG. The duct burner located between split HP superheater sections (Fig. 2.13) is most common. This location allows the HRSG designer to balance the amount of superheater and reheater surface areas upstream and downstream of the duct burner for steam temperature control. HRSGs have also been designed with the duct burner directly located upstream of the HP evaporator surface. For some cogeneration applications, two duct burners located in different sections of the HRSG have been used to increase both HP steam production and a lower-pressure steam flow rate. The amount of oxygen remaining downstream of the first duct burner limits the size of the second duct burner.
2.3.3.4 Stack temperature
Intuitively, the lower the HRSG stack temperature, the greater the amount of energy that has been recovered. The familiar equation to calculate the amount of heat transferred (or “recovered” in the case of HRGs if losses are ignored) is presented below:
Q is the amount of heat transferred;
m is the mass flow rate of the heat source;
cp is the specific heat of the heat source; and
(T1−T2) is the temperature difference of the heat source between two points in the flow path.
With T1 the temperature of the turbine exhaust gas entering the HRSG and T2 the exhaust gas temperature immediately downstream of the last heat transfer surface area, the lower the T2 temperature is, the greater the waste heat recovery in the HRSG. The practical lower limit for the HRSG stack temperature is 150°F. This can be achieved with the use of proper metallurgy for cold end heat transfer surface area (i.e., LP economizer; also known as “preheater” or “feedwater heater”). If the entire LP economizer is fabricated with carbon steel tubes, then the realistic lower limit for the HRSG stack temperature is approximately 175°F and the condensate temperature entering the LP economizer should be controlled to around 140°F to 150°F to prevent external corrosion.
2.4 Power cycle variations that use HRSGs
A major attribute of HRSGs is their versatility. HRSGs can recover heat from the very smallest gas turbines to the very largest. They can also be configured for a myriad of power cycle variations. A very widely used power cycle variation is cogeneration. Cogeneration, as the name implies, is the simultaneous generation of two different forms of energy, most often electricity and steam. HRSGs are brilliantly suited for cogeneration applications with their ability to generate steam at three different pressure levels. HRSGs can also be used for cogeneration applications requiring electricity and hot water. Another power cycle variation that uses HRSGs is steam power augmentation (PAG). In this cycle, a portion or in some cases the total amount of steam generated in the HRSG is routed to the gas turbine and injected into the engine upstream of the power turbine. This additional mass flow into the turbine yields additional power output, hence, the term “power augmentation.” More recent power cycle variations that use HRSGs are the integrated gasification combined cycle (IGCC)and the solar hybrid cycle. Let’s explore each one of these power cycle variants in more detail.
2.4.1 Cogeneration
Cogeneration plants, also known as combined heat and power plants, burst onto the power generation scene in a big way during the Public Utility Regulatory Policies Act (PURPA) years of the 1980s. Although in use prior to then, cogeneration plants proliferated as a result of the PURPA of 1978. This US federal law created the qualifying facility (QF), entitling the QF owner to sell electricity to the utility company at an avoided cost rate. In order to meet the requirements of PURPA, the cogeneration QF had to meet a certain efficiency threshold. This is where the HRSG came into play. By using the gas turbine’s exhaust energy, the HRSG produced steam and/or hot water, which could then be sent to another facility for beneficial use. The electricity generated from the gas turbine, and for many cogeneration QF plants, the additional electricity from a steam turbine, was then sold to the local utility at the utility company’s avoided cost rate. Although the PURPA laws have changed, cogeneration plants continue to be built to service hospitals, universities, food processors, refineries, and petrochemical facilities, to name a few industries benefitting from the efficiency of generating two forms of energy at the same time.
In its basic form, a cogeneration plant can consist of a gas turbine generator exhausting into a heat recovery steam generator, with the HRSG producing either steam or hot water as thermal energy. Fig. 2.15 depicts a cogeneration plant with a two-pressure level HRSG. The HRSG is producing HP steam and LP steam for process use.
Several successful enhanced oil recovery cogeneration plants have been constructed, where saturated steam produced in the HRSG is injected into an oil field to increase oil production rates. In this arrangement the HRSG is only producing steam at one pressure level.
The versatility of the HRSG makes configuring a cogeneration facility to meet the needs of the thermal host relatively easy since one, two, or three different steam pressures can be produced in a quite wide pressure range (25–2500 psig). Hot water can also be extracted from the HRSG for process use.
Another common adaption is the combined cycle cogeneration plant. In this power cycle variation, a combined cycle plant provides a portion of the steam produced in the HRSG for process use. With this cycle, not only do you get the high efficiency of the combined cycle, but also the added efficiency benefit of the process steam energy content.
The combined cycle cogeneration plant adds another layer of cycle configuration flexibility. The steam turbine can be a backpressure machine, a condensing machine, a condensing machine with a single extraction, or a condensing steam turbine with double extractions. IP and/or LP steam generated in the HRSG can either be admitted to the steam turbine or matched to a process steam pressure level for direct routing to the thermal host. Incorporating a duct burner into the HRSG provides even greater steam production flexibility to match the thermal host’s varying steam needs.
The following two figures illustrate the versatility of the combined cycle cogeneration plant. Fig. 2.16 contains a backpressure steam turbine exhausting to a high-pressure or medium-pressure (MP) process steam header. The LP steam generated in the HRSG is routed directly to the LP process steam header. Depending on the gas turbine used and the thermal host’s steam levels, the HRSG could also be fitted with an IP level, with the IP steam routed to the MP process steam header.
The combined cycle cogeneration plant shown in Fig. 2.17 is a bit more complex. The HRSG has three pressure levels and supplementary firing. The duct burner is nested within the HP superheater sections. HP steam from the HRSG is admitted to the steam turbine throttle. A controlled extraction port in the steam turbine supplies the thermal host’s MP process steam header. The HRSG IP steam is admitted to the steam turbine for power generation. LP steam from the HRSG can either be sent to the thermal host or admitted into the steam turbine depending on process steam flow requirements. This power cycle cogeneration configuration is suited for F-class gas turbines and larger. Depending on the size of the steam turbine and surface condenser, all or some fraction of the total steam produced in the HRSG can be admitted to the steam turbine for electricity production.
Most combined cycle cogeneration plants are nonreheat. However, if MP process steam flow rates are in the 200,000-pound-per-hour range or less, it is possible to employ a reheat cycle design to marginally improve overall efficiency (Fig. 2.18). With this power cycle variation, a portion of the cold reheat steam is sent to the thermal host’s MP process steam header. As more and more cold reheat steam is diverted to process, the efficiency gain due to reheat becomes less. Furthermore, too much cold reheat sent to process results in tube metal design temperatures that start to approach a dry reheater design. It is for these two reasons that the practical limit of cold reheat steam flow diverted to process is roughly 200,000 pounds per hour.
Without a doubt, the versatility of the HRSG has greatly contributed to the success of the modern day combined heat and power plant.
2.4.2 Steam power augmentation
Steam power augmentation, or “steam injection,” is a means of increasing power output of a gas turbine by injecting additional mass flow through the power turbine section of the engine. The additional mass flow results in an incremental gain in power output since turbine work is directly related to mass flow (see the previously discussed equation: 
where m is mass flow through the power turbine). The power augmentation steam is injected upstream of the turbine section either downstream of the combustors or into the combustion section. When steam is injected into the combustion of the gas turbine, it has the added benefit of reducing engine NOx formation primarily by reducing the combustion zone mean temperature. Steam power augmentation for gas turbines with dry low NOx combustors must have the steam injected downstream of the combustors.
Fig. 2.19 depicts the steam power augmentation cycle for a simple cycle application. The HRSG is the source of the power augmentation steam by capturing some of the waste heat from the gas turbine exhaust. The HRSG shown in Fig. 2.19 has a drum, but a once-through HRSG design can also be used for simple cycle power augmentation installations.
Most purpose-built power augmentation plants for simple cycle applications use smaller gas turbines (less than 50 MW) as the prime mover. There are commercial installations where once-through HRSGs have been back-fitted to F-class simple cycle gas turbine installations to boost power output. The HRSGs were designed such that they could be operated dry (no water in the HRSG pressure parts). This allows the simple cycle gas turbines to continue in operation and exhausting through the HRSG without steam power augmentation in-service.
Steam power augmentation can also be used in combined cycle power plants. When additional power output is desired, cold reheat steam can be diverted upstream of the HRSG and sent to the gas turbine for power augmentation steam. This reduces the hot reheat steam flow to the steam turbine so some bottoming cycle power output is lost, but the gain in gas turbine output from the steam power augmentation results in an overall incremental gain in plant net output. The incremental heat rate for the additional power output is in the range of 10,000 to 11,000 Btu/kWh (HHV).
Another variation of power augmentation for combined cycle power plants is referred to as “hybrid power augmentation.” In this variation, the HRSG is fitted with a duct burner that can generate more HP steam than the steam turbine can admit through the throttle valves. The excess HP steam is used as power augmentation steam in total or in combination with cold reheat steam. See Fig. 2.20 for an illustration of the hybrid power augmentation cycle. The incremental heat rate for the additional power output is in the range of 12,000–15,500 Btu/kWh (HHV) depending on the amount of HP steam used for power augmentation steam.
Steam power augmentation for simple cycle applications finds a niche where additional plant output is desired but for some reason the plant cannot be designed or built out to combined cycle. Steam power augmentation can also be designed into a combined cycle power plant where the power market is financially attractive for peaking power at incremental heat rates north of 15,000 Btu/kWh (HHV).
2.4.3 Integrated gasification combined cycle
Coal-fired power plants have long been a mainstay of power generation worldwide. Predominately, coal is combusted in pulverized form for electricity generation. As the need for greater efficiency materialized, coal-fired cycle design added additional regeneration (more feedwater heating), then single reheat, and in some cases double reheat. Boilers went from subcritical to supercritical, and now are being designed for ultrasupercritical conditions (in excess of 4000 psia). Even so, the most efficient coal-fired Rankine cycle cannot match the efficiency of a standard combined cycle power plant. Yet, what if the fuel cost advantages of coal and petcoke could be married to the cycle efficiency of combined cycle power plants with the added bonus of cleaner coal combustion and possibly CO2 capture? From this economic and environmental stimulus, the IGCC was formulated, developed, and brought to commercialization. And once again, the HRSG has a major role in this power cycle variant.
The major components of an IGCC plant are the gasifier; the gas clean-up equipment, which can include CO2 capture; the air separation unit; and the combined cycle equipment (gas turbine, HRSG, steam turbine, etc.). Oxygen from the air separation unit is mixed with coal in the gasifier to produce synthetic gas (syngas). The hot syngas undergoes cooling, sulfur and particulate removal, and if desired, CO2 removal. The cooling of the syngas is one area of integration between the gasification process and the combined cycle power plant. Feedwater can be sent to cool the syngas, and the saturated steam produced in the syngas cooling stage is then returned to the HRSG for superheating and power production in the bottoming cycle.
Another area of integration is with the gas turbine. The gasification process requires relatively pure oxygen. The compressed air feed to the air separation unit can come from a separate air compressor or a portion of the compressed air can be obtained from the compressor section of the gas turbine. Nitrogen from the air separation unit is piped to the gas turbine and combined with the remaining air from the compressor, then mixed with the syngas for combustion in the gas turbine’s combustors. The resultant gas turbine exhaust is materially different, with much higher concentration of nitrogen. The HRSG design can readily accommodate the different exhaust gas composition. Fig. 2.21 provides a simplified diagram of the integration between the gasification process and the combined cycle.
2.4.4 Solar hybrid
Since the mid-2000s, solar power has gained traction and is on the cusp of generating appreciable amounts of electricity as a percentage of total worldwide electrical consumption. At the present time, photovoltaic (PV) power dominates the solar power sector due to capital cost and its distributed nature. PV can be installed on carports, residential roofs, even office building exterior walls. Solar power can also take the form of CSP, where utility scale installations of mirrors (heliostats) concentrate solar radiation to a central tower. Within the tower a working fluid is heated, which in turn transfers heat to water for the generation of steam. The steam then drives a steam turbine generator in a conventional Rankine cycle. Another form of solar power is the solar hybrid power plant.
Solar hybrid is a more recent power cycle variant of combined cycle, where parabolic troughs or linear Fresnel collectors heat a working fluid (see Fig. 2.22 for the cycle diagram).
The hot working fluid is circulated through a steam generator, which transfers the heat to water thereby generating saturated steam. The saturated steam exits the solar steam generator and is sent to the HRSG, where it mixes with saturated steam exiting the HRSG’s HP drum. The combined saturated steam flow then flows to the HP superheater section of the HRSG, and once superheated, is sent to the steam turbine. The HP steam produced from the sun’s energy in the solar field; in a practical sense, replaces the HRSG duct burner generated HP steam. It does it though without burning additional fuel; hence, the overall cycle heat rate improves. This is in contrast to the negative impact on heat rate from the duct burner.
It is also possible to directly capture the solar radiation right to water thereby eliminating the heat transfer fluid loop. The steam generated in this fashion would also mix with the saturated steam generated in the HRSG.
2.5 Conclusion
Energy powers our modern lifestyle, from transportation, to the manufacture of goods, to keeping the lights on, to everyday tasks such as food storage and preparation. One form of energy, electricity—especially inexpensive electricity—is crucial for the world’s economy. It has been humankind’s quest for inexpensive electricity that has taken us from using the unique Rankine and Brayton cycles to generate electricity to the present day combination of these two distinct cycles into a “combined cycle.”
As we have discussed in this chapter, the HRSG is the bridge between the Brayton (gas turbine) and the Rankine (steam turbine) cycles to technically allow the combined cycle power plant. HRSGs take the high-temperature but low-pressure gas turbine exhaust and recover this energy to generate high-temperature steam at various pressure levels for power generation in the steam turbine. HRSGs can generate up to three different steam pressures as well as produce reheat steam for higher cycle efficiencies. Supplementary firing and emission control hardware can also be integrated into the HRSG design to generate additional steam and reduce gaseous emissions, respectively.
HRSGs are versatile. They can be used to recover energy from the exhaust gas on the smallest to the very largest gas turbine models. The versatility of HRSGs is also demonstrated in the variants of the combined cycle that use HRSGs. Combined heat and power plants (cogeneration plants), the power augmentation cycle, the IGCC, and the solar hybrid power plant all require the venerable HRSG to work efficiently and reliably.