2.2.1 Pool boiling

When a large volume of liquid is heated by a submerged heating surface and the motion is caused by free convection currents stimulated by agitation of the rising vapor bubbles, it is called pool boiling [2]. During the course of heating it is noted that a general correlation lies between the gradually increasing heat flux, q, from the immersion heater and the corresponding temperature difference between the surface of the heater and the bulk of liquid, ΔT. The result of such correlation is presented in Figure 2.3 [1].

At point A since the heat flux is very low, no boiling occurs and no bubbles are formed, hence movement of liquid is by natural convection. In the region A-B, phase change takes place only as evaporation at the free surface of liquid that is at the liquid-vapor interface, and the corresponding heat flux is proportional to ΔT^4/3 [2].

A further increase in heat flux will result in vapor bubble formation at the hot surface, yet the bulk of the liquid may still be below saturation temperature. The bubbles, which form in the hot boundary layer, condense as they rise into the colder liquid, giving up their latent heat to raise the temperature of the water. This phase change still occurs because of evaporation at the free surface. Bubble agitation in this region, B-C, is insignificant, since there is only a minor change in the slope of the curve. The formation and release of the steam bubbles at the metal water interface, with water still wetting the surface, cause ΔT to remain low. This region, wherein heat transfer takes place by natural convection, is called subcooled nucleate boiling [3].

Further rising of heat flux results in rapid formation of a large number of bubbles on the hot surface. As a result, strong local velocity is encountered within the liquid, causing a further increase in heat transfer. This is the region of nucleate boiling, where the bulk of the fluid is at saturation temperature and bubbles travel through bulk of the liquid to emerge at the free surface. Beyond point C bubbles do not collapse. This region, C-D, is one of stable boiling, and has a very steep slope. The heat flux is proportional to ΔT^m, where “m” lies between 2 and 4 [3]. In this region, as the surface temperature is increased, the heat flux also increases up to a maximum value at point D.

Nucleate boiling is incipient boiling, which is characterized by the growth of bubbles on the heated surface that rise from discrete points on a surface, whose temperature is only slightly above the liquid’s saturation temperature. In general, the number of nucleation sites is increased by an increase in surface temperature. An irregular surface of the boiling vessel (i.e., increased surface roughness) can create additional nucleation sites, while an exceptionally smooth surface, such as glass, lends itself to superheating. Under these conditions, a heated liquid may experience boiling delay and the temperature may go somewhat above the boiling point and fail to boil.

Beyond the nucleate boiling region, the bubbles of the steam forming on the hot surface begin to interfere with the flow of water to the surface and eventually coalesce to form a film of superheated steam over part or all of the heating surface. This thin layer or film of vapor has low thermal conductivity and thus insulates the surface and reduces heat transfer and the heat flux, which in turn results in higher metal temperature. When vapor film insulates the surface from the liquid it is called film boiling. Region D-E shows the “onset of film boiling.” At point E bubbles become so large and numerous that liquid has difficulty approaching the surface as the bubbles rise and thus starve the surface of liquid. Consequently there is a sudden jump in heating surface temperature, as shown at point E′. With water at atmospheric pressure, ΔT might be 25K at E and more than 1000K at E′ [2]. At this temperature there is a real danger of heating surface burning out or melting. Hence, point E is known as the burn out point. It is therefore imperative to operate as close as possible to point E that is at point D, without the risk of burn-out. If at point E a situation arises when heat flux would be dependent on the surface temperature, as when a fluid is condensing at variable pressure, it will be possible to reach point F from point E.

Region E-F in Figure 2.3 shows unstable film boiling where the vapor blanket alternately breaks down and recovers. As a result there is a decline in heat flux with increase in surface temperature.

Beyond point F, in the region F-G, the vapor film becomes thicker and film boiling becomes stable, where heat flux again increases with an increase in ΔT. Heat is transferred by conduction, convection and radiation across the layer of vapor that blankets the heating surface. At this region only evaporation takes place without formation of any bubble. This region thus is identified as stable film boiling one. If now the heat flux is reduced below the value at point G, the conditions do not jump back to E, but follow a path to point F. At this point the vapor film suddenly collapses and conditions revert to those at point H on the nucleate boiling curve.

Point D, which is very close to point E, is known as the point of departure from nucleate boiling (DNB) or the “critical heat flux.” Heat flux at D is extremely high, as for example with water at atmospheric pressure heat flux is about 1500 kW/m² [2]. The temperature difference, ΔT, between the boiling liquid and the heating surface at which the critical heat flux occurs, is known as the critical temperature difference.

During operation of the boiler it is difficult to theoretically predict the heat transfer coefficients of nucleate boiling that would ensure ΔT remains below the critical temperature difference. Nevertheless, the ability to predict the value of the maximum heat flux, q_max, i.e., point E in Figure 2.3, is also useful as this represents an upper limit in the nucleate boiling heat transfer. The following formula, from Roshenow and Griffith, gives the value of q_max in SI units, i.e., W/m² [3].

$q_{\max }/({\rho }_V*\lambda )=0.0121*{\{({\rho }_L-{\rho }_V)/{\rho }_V\}}^{0.6}$ (2.3)

(2.3)

where

Once the critical heat flux data is determined it is compared with the required heat flux data. For the design to be acceptable, the critical heat flux, which is dependent on the following parameters, must always be greater than the heat flux generated in the course of boiling.

2.2.2 Forced-convection boiling

Besides pool boiling, vapor is also generated by passing liquid through a tube heated either by firing fuel, as in a once-through boiler, or by condensing steam, as is commonly used in process industries. This method of generating vapor is known as forced-convection boiling.

The nature of boiling in forced convection is similar to pool boiling in many ways like the regimes of nucleate or film boiling. The most important difference is in the cause of “burn-out,” since the vapor and liquid must travel simultaneously through the tube. As a result both heat transfer and pressure drop are affected by the behavior of the two-phase flow, which keeps on changing along the tube due to gradual evaporation of liquid. In a once-through tube, liquid below or at the saturation point enters at the bottom and gradually evaporates until a dry or superheated vapor leaves at the top. In the process, some or all of the following regions are encountered by the flow system (Figure 2.4):

The first and the fourth regions are extremes consisting of liquid only and vapor only regions, respectively. In these regions in turbulent flow the phenomena of heat transfer to water and to superheated steam in pipes is described by the Dittus-Boelter equation as follows [3].

For turbulent flow through tubes the phenomenon of heat transfer to water is described by

$Nu=0.023*{(\mathit{Re})}^{0.8}{(\mathit{Pr})}^{0.4}$ (2.4)

(2.4)

For superheated steam through pipes, the equation is

$Nu=0.0133*{(\mathit{Re})}^{0.84}{(\mathit{Pr})}^{0.33}$ (2.5)

(2.5)

where

From Eq. 2.4, the heat transfer coefficient of water is found as

$h_{water}=1063*(1+2.93*{10}^{-3}*T_f)*(V_f^{0.8}/d^{0.2})$ (2.6)

(2.6)

From Eq. 2.5, the heat transfer coefficient of superheated steam is

$h_{steam}=4.07*(1+7.69*{10}^{-4}*T_g)*\{V_g^{0.79}/(d^{0.16}*l^{0.5})\}$ (2.7)

(2.7)

where

Example 2.1

In a once-through boiler, feedwater enters through the bottom of tubes, receives heat from combustion of coal, and escapes from the top of the tubes as superheated steam. Calculate the heat transfer coefficients of water and superheated steam using the following data:

Pressure of feedwater at tube inlet=27.0 MPa

Absolute temperature of feedwater, T_f=563 K

Pressure of superheated steam at tube outlet=24.5 MPa

Absolute Temperature of superheated steam, T_g=838 K

Velocity of Feedwater through each tube, V_f=6 m/s

Internal diameter of each tube=63.5 mm

Length of pipe=30 m

Solution: The density of feedwater at 27.0 MPa and 563 K is

${\rho }_f=763.942{\text{kg}/\mathrm{m}}^3$

The density of superheated steam at 24.5 MPa and 838 K is

${\rho }_g=74.532{\text{kg}/\mathrm{m}}^3$

For constant mass flow through each tube,

${\rho }_f*V_f*A={\rho }_g*V_g*A$

Hence, the velocity of superheated steam through each tube is

$V_g={\rho }_f*V_f/{\rho }_g=763.942*6/74.532=61.5\mathrm{m}/\mathrm{s}$

Applying Eq. 2.6, the heat transfer coefficient of water is

$h_f=1063*(1+2.93*{10}^{-3}*563)*(6^{0.8}/{0.0635}^{0.2})=20496.43{\mathrm{W}/\mathrm{m}}^2\mathrm{K}$

Applying Eq. 2.7, the heat transfer coefficient of superheated steam is

$h_g=4.07*(1+7.69*{10}^{-4}*838)*\{{61.5}^{0.79}/({0.0635}^{0.16}*{30}^{0.5})\}=49.18{\mathrm{W}/\mathrm{m}}^2\mathrm{K}$

2.2.3 DNB (departure from nucleate boiling)

At the onset of boiling, metal temperature remains marginally above the saturation temperature of fluid. This is accomplished by maintaining adequate supply of liquid to avoid the occurrence of DNB. As heating continues, the liquid phase changes to vapor and the DNB point (point D in Figure 2.3) is reached, which is also the end point of the nucleate boiling. The metal temperature starts increasing at this point until point E is reached. In the region E-F, the metal temperature decreases, but increases again in the superheat region beyond point F. In the event the applied heat flux exceeds the critical heat flux, which is a function of pressure, the DNB would occur at low steam quality causing the metal temperature to be high enough to melt the tube.

In an optimistic design, in most boilers, ΔT is of the order of 3–6 K, which is well below the critical temperature difference of water at point E, i.e., about 25 K at atmospheric pressure. Thus ensuring the design value of heat flux is much lower than the maximum value or the critical heat flux.

2.2.4 Circulation

In drum-type subcritical boilers, water flows from the drum through downcomers to the bottom of the furnace, then moves up through risers or evaporator tubes and returns to the drum. From Figure 2.5 it can be seen that the water in the downcomers does not receive any heat while risers, which form furnace walls, absorb the heat of combustion of the fuel. As a result, the water inside the riser tubes is heated and gradually steam bubbles form in the water. This mixture of water and steam in the riser tubes has a much lower density than the density of water alone in the downcomers. A density difference thus generated results in a static head difference causing thermo-siphon effect, which provides the driving force for a downward flow in downcomers and an upward flow in risers. Since the flow of fluid is generated by density difference alone, this is known as natural circulation in boilers [1,4,5].

For maintaining continuous circulation of fluid, the fluid flow must overcome friction losses in downcomer and riser tubes, headers, bends, etc. One way of overcoming resistance due to friction is by increasing heat input to the furnace. As the heat absorption increases the density difference between the fluid in the downcomers and the fluid in the risers also increases, resulting in an increase in fluid flow. However, an increase in fluid flow in turn also causes an increase in friction loss. As a result, a point is reached when the rate of increase in friction loss surpasses the benefit of the enhanced density difference, where the rate of fluid flow starts to “drop.” Hence, all natural circulation boilers are designed to operate below the point of “drop.” The phenomenon of rising flow-rising heat input tends to make a natural circulation boiler self-compensating in the case of sudden overloads, changes in heat-absorbing surface condition, non-uniform burner disposition, etc.

As the operating pressure of fluid is increased, the density difference between the saturated liquid and the saturated vapor starts diminishing and reduces to zero at critical pressure, as is evident from Figure 2.6. Thus it becomes imperative to raise the height of the boiler higher to maintain adequate circulation with higher-pressure boilers.

In practice, at 18 MPa or higher fluid pressure the density difference is so small that it is difficult to maintain natural circulation, even after substantially increasing the height of the boiler. In such cases the fluid flow in the boiler tubes is ensured with the help of assisted or forced circulation.

In assisted or forced circulation, pumps are incorporated in the downcomer circuit to help natural circulation overcome friction losses. Downcomer tubes from the boiler drum are connected to a common header that serves as the suction header of the pumps. Discharge piping of these pumps is then connected to another header from which riser tubes are issued. Since forced circulation can overcome higher-resistance circuits smaller diameter riser tubes, than the tube diameter of natural circulation boilers, are used to economize the cost of tube materials. To reap the economic benefits of assisted circulation, this type of circulation is also adopted in some boilers operating at a pressure below 18 MPa.

The process of circulation, be it natural or forced, is dominated by what is called the circulation ratio, which is defined as the ratio of the “quantity of water in circulation” to the “quantity of steam produced.” For a natural circulation boiler (up to 18 MPa) the circulation ratio may range from 4 to 30, while for forced circulation boiler the circulation ratio varies from 3 to 10 [6].

2.3.1 Water and steam circuit

At operating pressure, temperature of the feedwater, prior to entering the steam generator, is less than that at the saturated condition, i.e., the entering feedwater is under the subcooled condition. Within the steam generator, the feedwater temperature is elevated to the near-saturation temperature in the economizer. Thereafter, the sensible heat for bringing the feedwater to saturation and the latent heat for evaporation of saturated feedwater is added in the water-wall for generating steam. This steam is separated from the steam-water mixture coming out from the water-wall and purified in the boiler drum to ensure the supply of saturated steam. The saturated steam is further superheated to attain the desired energy level.

The heat-transfer surfaces discussed so far, e.g., economizer, water-wall, superheaters/reheaters, etc., are located in the steam generator in the flow path of products of combustion so that heat is absorbed efficiently in proper proportions in various heat transfer zones.

The evaporation zones are comprised of the boiler drum, downcomers, and riser tubes or simply risers. While the boiler drum and downcomers are located outside the heat transfer zone, the riser tubes are exposed to heat transfer. While flowing through the riser tubes the feedwater receives the heat of combustion for evaporation. The riser tubes in modern steam generators are arranged such that they form the enclosure of the combustion chamber or furnace and receive the heat. Hence, riser tubes are usually called water-walls.

The feedwater from the economizer enters the boiler drum, flows down through the downcomers, passes through the pipes to the water-wall bottom header and rises through the riser tubes. In the riser tubes, the feedwater absorbs the heat, part of which is converted to steam and re-enters the boiler drum as a steam-water mixture. From this mixture, steam is separated in the boiler drum and purified. Dry saturated steam from the boiler drum is lead out through the saturated steam pipes to the superheaters. The complete flow path of working fluid is shown in Figure 2.7.

The process of separation and purification of steam in the boiler drum is accomplished by drum internals, e.g., cyclones, baffles, etc., chemical and feedwater admission piping, blow-down lines, etc. (Figure 2.8). The process includes three steps: separation, steam washing, and scrubbing.

Separation is the process of removing the bulk mass of water from steam and is accomplished by any one of following means:

Steam washing, scrubbing, or steam drying is the process of passing steam between closely fitted corrugated plates or screens on which mists are deposited in a way similar to that occurs during filtration. To avoid re-entrainment of water, velocity is kept low. The collected water drips to the boiler water below by gravity.

The arrangement of drum internals must be such as to ensure the following [1,3]:

The function of the superheaters and reheaters is to raise the total heat content of the steam. There are two fundamental advantages of providing superheaters and reheaters: first, gain in efficiency of the thermodynamic cycle and second, enhancement of life in the last stage of the turbine by reduction of moisture content of the flowing steam.

In low-temperature boilers only convection superheaters are used, but large, high-temperature steam generators use both radiant and convection superheaters along with the reheaters to achieve the required superheat and reheat final steam temperatures. Today, modern steam power plants run at lower load at rated conditions during off-peak hours. Hence, the superheater/reheater elements are designed to achieve a normal operating temperature at about 60% of the boiler maximum continuous rating (BMCR) load for better steam turbine performance. The location of superheaters and reheaters in the flue-gas path are normally split into primary and secondary sections to facilitate reasonably unchanged steam temperature control.

2.3.2 Furnace

In the furnace of a steam generator combustion of fuel takes place with atmospheric air to release heat. Hence, it is essential to ensure complete combustion of fuel under all operating conditions to harness the maximum heat potential of the fuel under combustion. (Detailed treatment of this process is included in Chapter 3, Fuels and Combustion.)

2.3.3 Fuel-burning system

The primary function of a fuel-burning system is to provide controlled and efficient conversion of chemical energy of the fossil fuel into heat energy. The heat energy thus generated is transferred to the heat-adsorbing surfaces of the steam generator. The main equipment for firing fossil fuel in a furnace is the burner.

An ideal fuel-burning system in a furnace is characterized by the following:

2.3.4 Draft system

For sustaining a healthy combustion process in a furnace it needs to receive a steady flow of air and at the same time remove products of combustion from the furnace without any interruption. When only a chimney is used for the release of products of combustion, the system is called a natural draft system. When the chimney is augmented with forced draft (FD) fans and/or induced draft (ID) fans the system is called a mechanical draft system. Small steam generators use natural draft, while large steam generators need mechanical draft to move large volumes of air and gas against flow resistance. A furnace utilizing an FD fan only is subjected to “positive draft” (pressurized), while one using an ID fan only is subjected to “negative draft” (subatmospheric pressure). Normally large coal-fired steam generators utilize balanced draft, employing both FD fans and ID fans. Figure 2.9 presents the flow path of a typical balanced draft furnace.

2.3.5 Heat recovery system

The flue gas or products of combustion leaving the convection heat-adsorbing surfaces in large steam generators still carry considerable heat energy at a temperature much above the steam saturation temperature. This heat, if exhausted to atmosphere without further heat transfer, will result in considerable losses. As a natural consequence of economic operation of a steam generator, a major part of this heat is recovered by placing the economizer thereafter the air heater in the exit flue-gas path.

Theoretically it is possible to reduce the flue-gas temperature to that of the incoming air to the air heater, but for technical and economic considerations, as described below, such reduction in exit flue-gas temperature is prohibited.

In an economizer the incoming feedwater absorbs heat from the exiting flue gases, raising the temperature of the feedwater as close as possible but lower than its saturation temperature prior to entering into the evaporating circuit.

The remaining heat in flue gas downstream of the economizer is recovered in air heater to raise the temperature of combustion air before it enters the furnace. The energy thus recovered returns to the furnace and reduces the amount of heat that must be released by the fuel to maintain the desired furnace temperature, thereby ensuring reduction in the amount of fuel to be burned. The hot air is also used to dry coal in a pulverized coal-fired boiler.

2.4.1 Fire-tube boiler

In this type of boiler hot gases pass through the tubes and water is contained in the shell. As the name suggests, its general construction is as a tank of water perforated by tubes that carry hot flue gases from the fire. The tank is usually cylindrical in shape to realize the maximum strength from simple structural geometry. The tank may be installed either horizontally or vertically.

Typically the furnace and the grate, on which fuel is burnt, are located underneath the front end of the shell (Figure 2.10). The gases pass horizontally to the rear, then either release through the stack at the rear or reverse directions and pass through the horizontal tubes to stack at the front. The fire tubes can be placed horizontally or vertically or at an inclined plane in a furnace.

In this type of boiler boiling takes place in the same compartment where water is stored. As a result, only saturated steam used to be produced in older designs of this type of boiler, but today, a fire-tube boiler may generate superheated steam as well.

A fire-tube boiler is sometimes called a “smoke-tube boiler” or “shell boiler.” This type of boiler was used on virtually all steam locomotives in horizontal “locomotive” form. It has a cylindrical barrel containing the fire tubes, but also has an extension at one end to house the “firebox.” This firebox has an open base to provide a large grate area and often extends beyond the cylindrical barrel to form a rectangular or tapered enclosure.

Fire-tube boilers are also typical of early marine applications and small vessels. Today, they are used extensively in the stationary engineering field, typically for low-pressure steam use such as for heating a building. However, the steam-generating capacity and outlet-steam pressure of these boilers are limited, therefore, they are unable to meet the needs of larger units. Another disadvantage of these boilers is that they are susceptible to explosions.

2.4.2 Water-Tube Boiler (Figure 2.11)

To generate high evaporation rate accompanied by high steam pressure, the fire-tube boiler becomes exorbitantly heavy; therefore, the size and weight become extremely difficult to manage. In regards to the size of the shell, these shortcomings are circumvented by passing flue gases outside the tubes, instead of inside, and water is circulated through the tubes for evaporation. Baffles are installed across the tubes to allow cross flow of flue gases to ensure maximum exposure of the tubes. On the basis of configuration of tubes inside the furnace, this boiler is further classified as “straight-tube boiler” and “bent-tube boiler.”

In a water-tube boiler water circulates in tubes heated externally by the hot flue gas. Fuel is burned inside the furnace, creating hot gas that heats up the water in the steam-generating tubes. Cool water at the bottom of the steam drum returns to the feedwater drum of small boilers via large-bore “downcomer tubes,” where it helps pre-heat the feedwater supply. In large utility boilers, feedwater is supplied to the steam drum and the downcomers supply water to the bottom of the water-walls. The heated water then rises into the steam drum. Here, saturated steam is drawn off the top of the drum. In large utility boilers water-filled tubes form the walls of the furnace to generate steam and saturated steam coming out of the boiler drum re-enter the furnace through a superheater to become superheated. The superheated steam is used in driving turbines.

2.4.3 Natural circulation boiler

Boilers in which motion of the working fluid in the evaporator is caused by thermo-siphon effect on heating the tubes are called “natural circulation boilers.” As discussed in section 2.2 circulation of water in natural circulation boilers depends on the difference between the density of a descending body of relatively cool and steam-free water and an ascending mixture of hot water and steam. The difference in density occurs because the water expands as it is heated, and thus, becomes less dense. All natural circulation boilers are drum-type boilers.

2.4.4 Forced/assisted circulation boiler

As noted in section 2.2 the density difference between the saturated liquid and saturated vapor starts diminishing at 18 MPa or higher fluid pressure, thus it is difficult to maintain natural circulation of fluid flow in boiler tubes. In such cases fluid flow is ensured with the help of forced/assisted circulation using pumps. The forced/assisted circulation principle applies equally in both supercritical and subcritical ranges.

2.4.5 Subcritical boiler and supercritical boiler

The critical pressure is the vapor pressure of a fluid at the critical temperature above which distinct liquid and gas phases do not exist. As the critical temperature is approached, the properties of the gas and liquid phases become the same, resulting in only one phase. The point at which the critical temperature and critical pressure is met is called the critical point. The critical pressure and critical temperature of water and steam are 22.12 MPa and 647.14 K, respectively. Any boiler that operates below the critical point is called a subcritical boiler, and one that operates above the critical point is known as a supercritical boiler. (Supercritical boilers are covered in section 2.10.)

2.4.6 Drum-type boiler

In a drum-type boiler the drum acts as a reservoir for the working fluid. These boilers have one or more water drums, depending on the size and steam-generating capacity. The drum is connected to cold downcomers and hot riser tubes through which circulation of working media takes place. The lower portion of the drum with feedwater is called the water space and the upper portion of the drum occupied by steam is called the steam space. A drum-type boiler can be either the natural circulation type or forced/assisted circulation type. Drum-type boilers are essentially subcritical boilers; they operate below the critical pressure of the working fluid. The economic design pressure limit of fluid in a drum-type boiler is around 18 MPa.

2.4.7 Once-through boiler

This type of boiler does not have a drum. Simply put, a once-through boiler is merely a length of tube through which water is pumped, heat is applied, and the water is converted into steam. In actual practice, the single tube is replaced by numerous small tubes arranged to provide effective heat transfer, similar to the arrangement in a drum-type boiler. The fundamental difference lies in the heat-absorbing circuit. Feedwater in this type of boiler enters the bottom of each tube and discharges as steam from the top of the tube. The working fluid passes through each tube only once and water is continuously converted to steam. As a result there is no distinct boundary between the economizing, evaporating, and superheating zones. The circulation ratio of this type boiler is unity [6]. These boilers can be operated either at subcritical or at supercritical pressures.

2.4.8 Stoker-fired boiler

Prior to the commercial use of “fluidized-bed boilers” (Chapter 5), stoker-fired boilers provided the most economical method for burning coal in almost all industrial boilers rated less than 28 kg/s (100 tph) of steam. This type of boiler was capable of burning a wide range of coals, from bituminous to lignite, as well as byproducts of waste fuels. However, over the years this type of boiler has become less popular due to more efficient technological advancement.

In stoker-fired boilers coal is pushed, dropped, or thrown on to a grate to form a fuel-bed. Stokers are divided into two general classes: overfeed, in which fuel is fed from above, and underfeed, wherein fuel is fed from below. Under the active fuel-bed there is a layer of fuel ash, which along with air flow through the grate keeps metal parts at allowable operating temperatures. Figure 2.12 and Figure 2.13 show a chain-grate overfeed stoker and a traveling grate overfeed stoker, respectively.

A comparison of the traveling grate stoker and the spreader stoker is given in Table 2.1.

2.4.9 Fossil fuel-fired boiler

Coal, fuel oil, and natural gas are the main types of fossil fuel [8]. These fuels can generate a substantial quantity of heat by reacting with oxygen. These fuels consist of a large number of complex compounds comprised of five principal elements: carbon (C), hydrogen (H), oxygen (O), sulfur (S), and nitrogen (N). Any of these three fuels, i.e., coal, fuel oil, and natural gas, can be used in steam power stations, but coal plays a large role in power generation because of the enormous reserves of coal in many countries around the world, e.g., the United States, China, Mongolia, India, Indonesia, Russia, Poland, South Africa, Australia, etc. Most of the large steam-generating stations in these countries are coal based, and coal is expected to remain a dominant fuel for this purpose for many years. Coal-fired power plants at present account for about 41% of power produced globally [9].

Coal-based steam-generating stations also dominate in those countries where natural gas and fuel oil are abundantly available, since the cost of natural gas and fuel oil is exorbitantly high compared to the cost of coal. Additionally, natural gas and fuel oil are more economical in other industries. In modern, large-capacity, coal-fired steam generators coal is burnt in suspension. (Pulverized coal-fired boilers are covered in more depth in Chapter 4.)

2.4.10 Fluidized-bed boiler

Fluidized-bed combustion ensures burning of solid fuel in suspension, in a hot inert solid-bed material of sand, limestone, refractory, or ash, with high heat transfer to the furnace and low combustion temperatures (1075–1225 K). The combustor-bed material consists of only 3–5% coal. Fluidized-bed combustion is comprised of a mixture of particles suspended in an upwardly flowing gas stream, the combination of which exhibits fluid-like properties.

Fluidized-bed combustors are capable of firing a wide range of solid fuels with varying heating value, ash content, and moisture content. In this type of boiler, pollutants in products of combustion are reduced concurrently with combustion – much of the ash and hence the particulate matter as well as sulfur is removed during the combustion process. Further, lower temperature combustion in the fluidized bed results in lower production of NO_X and obviates any slagging problem. (Fluidized-bed boilers are covered in Chapter 5.)

2.4.11 Waste-heat recovery boiler

A waste-heat recovery boiler (WHRB) or a heat-recovery steam generator (HRSG) is a heat exchanger that recovers heat from a gas stream and in turn produces steam that can be used in a process or to drive steam turbines. A common application for HRSGs is in a combined cycle power plant (CCPP), where hot exhaust gas from a gas turbine is fed to the HRSG to generate steam. (These types of boilers are discussed in Chapter 7.)

2.6.1 Convection superheater/reheater

Convection superheater/reheater heating surfaces are placed above or behind the banks of water tubes at the rear of the radiant heating surfaces to protect them from combustion flames and high temperatures. The main mode of heat transfer between the flue gases and the convection superheater/reheater tubes is convection. With an increase in steam demand, both fuel and air flow increase, resulting in an increase in flue-gas flow. As a result, convective heat transfer coefficients increase both inside and outside the tubes. The overall heat transfer coefficient between gas and steam hence increases faster than the increase in mass flow rate of steam. Thus, steam receives greater heat transfer per unit mass flow rate and the steam temperature increases with the load as shown in Figure 2.14.

2.6.2 Radiant superheater/reheater

Some sections of superheaters and reheaters are placed closer to the higher-temperature zone of the combustion flames for greater heat absorption. With this placement of the superheater/reheater sections, the heat transfer between the hot flue gases and the flame and the outer surface of the tubes takes place by radiation. Hence, these heating surfaces are called “radiant superheater/reheaters” [14].

From fundamental heat transfer theory it is known that radiation heat transfer is proportional to ($T_{\mathrm{flame}}^4–T_{\mathrm{tube}}^4$), where T_flame is the flame absolute temperature and T_tube is the tube surface absolute temperature. However, T_flame is much greater than T_tube and is also not dependent on load. As a result, the heat transfer is distinctly proportional to $T_{\mathrm{flame}}^4$ and the resulting heat transfer per unit mass of flow decreases as the steam flow increases. Hence, an increase in steam flow due to enhanced load demand would cause the final steam temperature to fall, which is opposite of that experienced by the convection superheater/reheater (Figure 2.14). To allow radiation through radiant heating surfaces they are arranged in flat panels or platen sections with wide spacing.

2.7.1 Counter-flow economizer (Figure 2.18)

It is always best for flue gases from the boiler to flow down across the economizer tubes and for water to enter at the bottom and flow up through the tubes, thus reducing both surface and draft losses. Up-flow of water eliminates unstable water flow, gives the most uniform gas distribution, and makes the proper use of a steaming economizer possible.

2.7.2 Steaming economizer

A steaming economizer may be defined as an economizer in which the temperature rise in the economizer is equal to or more than two-thirds of the difference between the economizer inlet feedwater temperature and the saturation temperature of the fluid at that pressure. This may be expressed mathematically as (T_O−T_I)=(2/3)*(T_S−T_I), where T_O is the economizer outlet feedwater temperature, T_I the economizer inlet feedwater temperature, and T_S is the saturation temperature of feedwater at its prevailing pressure. The steam content of water at the economizer outlet should not exceed 20% of the feedwater flow at full load, and should be lower at part loads.

When feedwater make-up is low, it is advantageous to produce part of the steam in an economizer rather than in the evaporating section. From commercial point of view steaming is not allowed in the economizer, since there is the possibility of tube failure due to two phase flows through the tubes.

2.7.3 Plain-tube economizer

In some designs of boilers a continuous loop-type economizer is used. Tubes of this type of economizer have relatively small diameter to escalate the magnitude of heat transfer and to minimize fouling.

2.7.4 Finned-tube Economizer (Figure 2.19)

Heat absorption in the economizer tubes may also be intensified by welding fins to the plain tubes on the flue-gas side. This type of arrangement helps make the economizer compact, and it also increases the use of metal per unit heating surface area, providing much more gain in heat absorption and requiring smaller space. These types of economizers are more suited to gaseous and liquid fuel-fired boilers and in combined cycle plants. In coal-fired boilers, finned tubes often get plugged with ash and are usually not used.

2.7.5 Return-bend tube economizer

When it is essential to clean the inside of the tubes mechanically this design is best. This type of economizer is basically a horizontal steel-tube economizer, one end of which is bent 180° and the other end is equipped with a special flange bolted with a return-bend fitting.

2.7.6 Staggered-tube economizer (Figure 2.20)

A staggered arrangement of economizer tubes is preferable for gas and oil-fired boilers. This arrangement provides ascending motion of the feedwater in the economizer, resulting in free release of flue gases entering from the top of the economizer tubes and flowing down across the staggered tubes leaving at the bottom. This economizer is not suitable for coal-fired boilers. In a coal-fired boiler tubes must be in line to allow free flow of ash chunks through the bank of tubes.

2.8.1 Recuperative air heater (Figure 2.21)

In a recuperative air heater, heat from the flue gases is transferred continuously to air through a heating surface. In this type of air heater the metal parts are stationary and heat is transferred by conduction through the metal wall. A commonly used recuperative air heater is the tubular type, but for lower air and flue-gas pressures the air heater can also be the plate type.

Tubular air heaters are normally the counter-flow shell-and-tube type where hot flue gases pass through the inside of the tubes and air flows outside. Baffles are provided to ensure maximum air contact with hot tubes. These heaters, however, consume much metal and occupy a large space.

2.8.2 Regenerative air heater

In a regenerative air heater, the heating surface is swept alternately by the flue gas and the air undergoes alternate heating and cooling cycles and transferring of heat by thermal storage capacity of the heat transfer surfaces. The gas and air counter-current flows move through the rotor separately and continuously. Thus, when half of the heating elements are exposed to flue gases at any instant, the other half is exposed to air. This air heater may include either rotating heat transfer surfaces with stationary air/gas distribution hoods (Ljungstrom design: Figure 2.22) or rotating air/gas distribution hoods, where the heat transfer surfaces are stationary (Rothemuhle design: Figure 2.23).

The rotor of a Ljungstrom air heater is driven by a motor and has several radial members that form sectors. Two opposite sectors are provided with stationary seal covers. The heat transfer surfaces are comprised of corrugated, notched, or undulated ribbing steel sheets that form baskets. They constitute the heat-storage medium of the air heater. As the rotating sectors enter the hot gas zone they are heated by the gas, and they store the heat as sensible heat. When they enter the air zone, they progressively release this heat to the air.

In a Rothemuhle air heater the storage elements are stationary and the gas and air ducts are connected to two rotating segments each [1]. The principle of heat transfer from hot gas to cold air is identical to that of the Ljungstrom air heater. Ljungstrom air heaters, however, are better for commercial use due to better heat transfer and lower leakage, sometimes less than 5%.

While they provide substantial fuel savings, air heaters are prone to common problems as discussed in the following.

2.8.2.1 Leakage

Leakage is a normal phenomenon in all air heaters. While leakage in a new recuperative air heater is essentially zero, leakage occurs with the passage of time and due to accumulation of thermal cycles. Nevertheless, leakage in a recuperative air heater can be restricted to 3% with regular maintenance.

With a rotary regenerative air heater air leakage is inherent and its design value ranges from 5–10%, which increases further over time as seals wear. When the higher-pressure air side passes to the lower-pressure gas side leakage occurs through the gaps between rotating and stationary parts.

The rate of such leakage is given by the following expression [1]:

$W_L=KA{(2g\mathrm{\Delta }P\rho )}^{0.5}$ (2.8)

(2.8)

where

W_L=Leakage flow rate, kg/s

K=Discharge coefficient (generally 0.4–1.0)

A=Flow area, m²

g=Acceleration due to gravity (1 kgm/Ns²)

ΔP=Pressure differential across gap, kg/m²

ρ=Density of leaking air, kg/m³

Ambient air infiltration through holes, doors, gaskets, expansion joints, etc., into lower-pressure gas streams is another form of air-heater leakage. The air-heater leakage can be determined by analyzing the %O₂ content of flue gases entering and leaving the air heater (dry basis) and is given by

$\%Leakage=\frac{\%{\mathrm{O}}_{2}Leaving-\%{\mathrm{O}}_{2}Entering}{21-\%{\mathrm{O}}_{2}Leaving}*0.9*100\text{[1]}$ (2.9)

(2.9)

Example 2.2

While evaluating the performance of a boiler it is observed from flue-gas analysis that %O₂ content of flue gases entering and leaving a rotary air heater are 5.0% and 7.5%, respectively. Determine the percentage infiltration of ambient air.

Solution: Applying Eq. 2.9, we get

$\%\mathrm{Leakage}=(7.5–5.0)/(21.0–7.5)*90=16.67\%$

Hence, the percentage of infiltration of ambient air in the boiler is 16.67%.

(NOTE: In a rotary air heater, whose seals are in good condition, leakage of air should not exceed 10%. The above result reveals that the seals of the air heater are in bad shape and replacement and proper adjustment of seals should be done.)

2.8.2.2 Corrosion

The air heater operates in the lowest temperature zone of the flue-gas path; as a result, the cold end of the air heater may be subjected to a temperature equal to or less than the dew point of the flue gases, resulting in a moisture film at the cold end. The formation of moisture in the cold end may cause corrosion and fouling. These adverse effects are exacerbated in units firing sulfur-bearing fuels. Upon combustion of the fuel, its sulfur content is oxidized to form SO₂, a portion of which is converted to SO₃ that combines with the moisture to form sulfuric acid vapor. This vapor condenses on surfaces at temperatures below its dew point of 393–423 K, resulting in potential corrosion in the cold end of the air heater.

The obvious solution would be to operate at cold-end metal temperatures above the acid dew point, but this results in unacceptable boiler heat losses. In practice, air heaters are designed to operate the same at minimum metal temperature (MMT), therefore, the resulting increase in boiler efficiency offsets the additional maintenance cost needed to prevent low-temperature gas corrosion. Some typical recommendations of leading boiler manufacturers regarding limiting MMTs while burning sulfur-bearing fuels are given in Figure 2.24 and Figure 2.25.

A generally accepted method of preventing gas corrosion is to raise the inlet temperature of the air in the steam coil air heater. Another method of overcoming corrosion is to apply corrosion-resistant coatings to low-temperature elements of an air heater. Cold-air bypass of air heaters or hot-air recirculation to combustion-air inlet to air heaters are other methods used to prevent gas corrosion.

2.9.1 Thermal insulation materials

There are four types of thermal insulation materials: granular, fibrous, cellular, and reflective. Granular materials, such as calcium silicate, contain air entrained in the matrix. Fibrous materials, such as mineral wool, contain air between fibers. Cellular materials, e.g., cellular glass and foamed plastics, contain small air or gas cells sealed or partly sealed from each other. Reflective insulation materials consist of numerous layers of spaced thin-sheet material of low emissivity, such as aluminum foil, stainless-steel foil, etc. In practice, a combination of two or more of the above four types are used.

2.9.1.2 Design of thermal insulation

From basic heat transfer theory [13] it is known that under steady state conditions the heat loss from a hot flat surface through thermal insulation to the ambient air is expressed as

$Q=(T_h-T_c)/(L/k)$ (2.10)

(2.10)

or

$Q=(T_c-T_a)*f$ (2.11)

(2.11)

Combining Eq. 2.10 and Eq. 2.11 we get

$Q=(T_h-T_a)/(L/k+1/f)$ (2.12)

(2.12)

where

Q=Heat loss, W/m²

T_h=Hot surface temperature, K

T_c=Cold surface temperature, K

T_a=Ambient air temperature, K

L=Thickness of insulation, m

k=Thermal conductivity of insulation material, W/m.K

f=Surface coefficient of insulation material, W/m².K

(combined effect of radiation and convection heat transfer)

The temperature of a cold surface or the surface of an insulation material can then be calculated from Eq. 2.11 as

$T_c=(Q/f)+T_a$ (2.13)

(2.13)

For steady flow of heat through composite walls, e.g., wall of a vessel lined with thermal insulation, Eq. 2.12 takes the shape of

$Q=(T_h-T_a)/(L1/k1+L2/k2+1/f)$ (2.14)

(2.14)

where

Q=Heat loss, W/m²

T_h=Hot surface temperature, K

T_a=Ambient air temperature, K

L₁=Thickness of wall, m

k₁=Thermal conductivity of material of wall, W/m.K

L₂=Thickness of insulation, m

k₂=Thermal conductivity of insulation material, W/m.K

f=Surface coefficient of insulation material, W/m².K

The expression of cold-surface temperature, T_c, remains the same as Eq. 2.13.

The heat loss from insulated cylindrical surfaces (e.g., pipes, tubes, small diameter vessels) can be written as

$Q=(T_h-T_a)/\left[\right\{r_i*\ln (r_o/r_i)*(1/k)\}(r_i/r_o)*1/f]$ (2.15)

(2.15)

where

Q=Heat loss, W/m²

T_h=Hot surface temperature, K

T_a=Ambient air temperature, K

r_i=Inner radius of insulation, m

r_o=Outer radius of insulation, m

k=Thermal conductivity of insulation material, W/m.K

f=Surface coefficient of insulation material, W/m².K

The temperature of cold surface or the surface of insulation material can then be calculated from

$T_c=(Q/f)*(r_i/r_o)+T_a$ (2.16)

(2.16)

For steady flow of heat through composite cylindrical surfaces, Eq. 2.15 changes to

$Q=(T_h-T_a)/\left[\right\{(r_i/k_1)*\ln (r_c/r_i)\}+\{(r_c/k_2)*\ln (r_o/r_c)\}+(r_i/r_o)*1/f]$ (2.17)

(2.17)

where

Q=Heat loss, W/m²

T_h=Hot surface temperature, K

T_a=Ambient air temperature, K

r_i=Inner radius of wall, m

r_c=Outer radius of wall/Inner radius of insulation, m

r_o=Outer radius of insulation, m

k₁=Thermal conductivity of material of wall, W/m.K

k₂=Thermal conductivity of insulation material, W/m.K

f=Surface coefficient, W/m².K

The expression of the cold-surface temperature remains the same as Eq. 2.16.

Example 2.3

Hot-surface temperature of a 150-mm thick fire brick wall is maintained at 1088 K. If the average thermal conductivity of the wall material is 0.196 W/m.K determine the maximum heat loss through the wall to ensure its cold-surface temperature does not exceed 333 K. For an ambient air temperature of 294 K, what is the surface coefficient of the wall material?

Solution: Applying Eq. 2.10, the maximum heat loss through the brick wall may be determined as

$\mathrm{Q}=(1088–333)/(0.15/0.196)=986.53{\mathrm{W}/\mathrm{m}}^2$

The surface coefficient of the wall material can be found from Eq. 2.11 as

$\mathrm{f}=986.53/(333–294)=25.296{\mathrm{W}/\mathrm{m}}^2\mathrm{.K}$

Example 2.4

Steam at a temperature of 623 K is flowing through a pipe with an outer diameter of 100 mm. The pipe is insulated with 20-mm thick mineral wool, the average thermal conductivity of which is 0.055 W/m.K. Determine the outside surface temperature of mineral wool if the ambient air temperature is 300 K. Assume the hot-surface temperature of the mineral wool is the same as that of the flowing steam and the value of the surface coefficient of the insulating material is 30 W/m².K.

Solution: The inside diameter of the insulating material may be assumed to be the same as the outside diameter of the pipe. Hence, r_i=50 mm and r_o=70 mm. Applying Eq. 2.15 we get

$\begin{array}{ll}\mathrm{Q}\hfill & =(623–300)/\left[\right\{(0.05/0.055)*\ln (0.07/0.05)\}+(0.05/0.07)\times 1/30]\hfill \\ \hfill & =323/(0.9091\times 0.3365+0.0238)\hfill \\ \hfill & =979.64{\mathrm{W}/\mathrm{m}}^2\hfill \end{array}$

The cold-surface temperature of insulating material is calculated from Eq. 2.16 as

$\begin{array}{ll}{T}_c\hfill & =(979.64/30\times 0.05/0.07)+300\hfill \\ \hfill & =323.32\mathrm{K}\hfill \end{array}.$

Example 2.5

One vessel having a carbon-steel wall of thickness 10 mm is carrying saturated steam and water at 441 K. The vessel is insulated with magnesia of thickness 50 mm. If the ambient air temperature is 303 K, determine the heat loss from the vessel.

Given:

i. thermal conductivity of carbon steel is 52 W/m.K

ii. thermal conductivity of magnesia is 0.05 W/m.K

iii. surface coefficient of insulation surface is 10 W/m².K

Solution: Assuming no temperature drop across the metal and applying Eq. 2.14 we get

$\begin{array}{ll}\mathrm{Q}\hfill & =(441-303)/(0.01/52+0.05/0.05+1/10)\hfill \\ \hfill & =151.83{\mathrm{W}/\mathrm{m}}^2\hfill \end{array}$

2.10.1 Introduction

From Carnot cycle we know that the efficiency of a thermodynamic cycle depends on the temperature of the heat source and the temperature of the heat sink and is independent of the type of working fluid. The higher the temperature of the heat source the lower the temperature of the heat sink and the more efficient the cycle. Since a temperature below 288 K for the heat sink and cooling water is rarely available, the efficiency can be increased by raising the temperature of the heat source. This gain in efficiency is substantially raised by using fossil fuel-fired plants in the thermodynamic supercritical zone.

The supercritical condition of a steam-water cycle is a state at which its temperature and pressure are above its thermodynamic critical point, where the pressure of the steam water is 22.12 MPa, the temperature is 647.14 K, and the density is 324 kg/m³. At the critical point, liquid-vapor phase shows the following special phenomenon:

As heat is applied to water above the critical pressure, the temperature rises, water molecules are gradually agitated, inter-molecular space increases uniformly, and fluid becomes less dense. Thus, at the critical point the liquid and gas phases of any fluid become a single supercritical phase. The transition from the dense-phase water with a compact molecular arrangement to a wide-spaced random arrangement of vapor is uniform. No internal bubbles are formed. Specific enthalpy changes uniformly and all other physical properties change continuously from liquid to vapor state with a gradual rise in temperature. Near to the critical point, small changes in pressure or temperature result in large changes in density. Thus, the density difference between the liquid and vapor phases reduces sharply.

The benefit of adopting supercritical technology in fossil fuel-fired power plants is not restricted to improvement in generating efficiency alone. Efficient operation of supercritical power plants leads to burning less fuel and consequently less SO_X, NO_X, and CO₂ emissions, which would in turn reduce global warming as well. The relative improvement in efficiency that can be achieved with a supercritical plant compared to a subcritical plant is presented in Table 2.2.

2.10.2 Steam-water circuit

Supercritical boilers do not have boiler drums. This type of boiler consists of a number of parallel circuits connected by inlet and outlet headers. Pressurized water enters the circuit at one end and leaves as superheated steam at the other end (Figure 2.26). As water absorbs energy it gradually expands in volume without internal bubbling and acts as extremely dense gas. Feedwater enters the bottom header of the economizer and passes upward through the economizer elements and then enters the evaporator, located downstream of the economizer.

In a supercritical boiler the evaporator circuit is the once-through type, but this type of boiler starts operating in the once-through mode beyond a particular minimum load of about 30–40%. Therefore, regardless of load it is necessary to ensure adequate flow of fluid within individual tubes so that the wall temperature does not exceed the allowable metal temperature limit. To meet this requirement below the minimum load, the supercritical boiler operates in circulation mode and needs a separator and circulation system for water-steam separation. Fluid coming from the evaporator enters the separator vessel, and the separated water is circulated back to the boiler.

There are two types of circulation systems in use today. In the first type, the separated water from the separator is led to the deaerator and is circulated to the feedwater system through the boiler feed pump. This system is simple and relatively inexpensive but involves loss of heat. In the other type of system a circulation pump is provided to circulate the water from the separator directly to the economizer (Figure 2.27). This prevents heat loss but increases cost.

From the separator vessel all fluid is routed through the pipes to the convection enclosure, similar to a subcritical design. There may be two to three stages of superheaters with interstage desuperheater for temperature control.

Another way of ensuring each tube of evaporator receives uniform total heat input above the minimum load is to deploy either inclined smooth tubes spiraled around the furnace enclosure (Figure 2.28) or a vertical wall furnace with rifled tubes (Figure 2.29).

Spiral construction of tubes provides the following benefits:

However, spiral-tube construction involves a complex support structure, which raises fabrication costs, and it is also relatively difficult to erect and repair spiral tubes.

The vertical water-wall design uses rifled tubing for improved mass flow without DNB or dry-out at part loads. Its advantages lie in ease of maintenance, along with the potential to lower costs, and it is much easier to erect, support, and repair. The main disadvantage of the rifled tube water-wall design is the overall height of the boiler is excessive.

2.10.3 Steam-water cycle

A typical steam-water cycle of a supercritical plant is shown in Figure 2.31, which shows that the expansion of steam from state 1 to the condenser pressure would cause the exhaust vapor to become very wet and may cause severe damage to the low-pressure turbine stages. Hence, this cycle has to essentially use single reheat, and if the high-pressure turbine inlet temperature is low (813 K or less) and pressure is high (above 26 MPa) double reheat may also be required.

2.10.4 Special features

2.10.5 Advantages

2.10.6 Disadvantages

Nonetheless, power plants operating with supercritical steam conditions show lower emissions of CO₂, SO_X and NO_X, higher efficiency, flexibility and operating reliability, reduced fuel costs, and investment costs compared to subcritical plants.

2.10.7 History

The first application of supercritical technology in a power plant took place in 1957 in the United States. It was the Philo Unit 6 of the American Electric Power with a capacity of 125 MW designed with the superheater outlet steam conditions of 31 MPa and 883 K along with a double-reheat steam temperature of 839/811 K. Within two years, the 325 MW Eddystone Unit 1 of the Philadelphia Electric Company started operating with superheater outlet steam conditions of 34 MPa and 922 K along with a double-reheat steam temperature of 839/839 K.

The performance of these plants, however, was unreliable, expensive, and they had poor operational flexibility, mostly related to thermal effects. As a result, most countries stayed with subcritical technology. However, development continued in the 1970s and 1980s in inland Europe and Japan to explore the benefits of supercritical technology.

It should also be noted that the former USSR was not far behind the United States in the use of supercritical technology. The first supercritical power plant with a capacity of 300 MW was commissioned in the USSR in 1963, and the first 800 MW unit was put into operation in 1968. Unlike the United States, the USSR approach toward adopting supercritical steam conditions was quite modest, while steam pressure was raised to 25.5 MPa, superheat steam temperature was kept the same as that of the subcritical plants, i.e., 818 K, and the reheat steam temperature was either the same as the superheat temperature or slightly higher, at 818–838 K. The world’s largest supercritical unit with a capacity of 1200 MW has been operating in the USSR since 1980 [16].

From the 1980s on, the technology got so matured that currently there are more than 600 supercritical power plants operating in the United States, Russia, Denmark, Germany, Poland, Japan, China, Korea, South Africa, India, Australia, and Taiwan. Power producers in other countries are also embarking on supercritical technology.

When considering supercritical steam conditions research shows that the industry is predominantly adopting a superheat steam pressure in the range of 24–31 MPa, with superheat/reheat steam temperatures in the range of 838–883 K [17]. In the near future the focus will likely be on adopting steam conditions of 31–32 MPa and 893 K/893 K with the ultimate goal of reaching ultra-supercritical steam conditions of 34–36 MPa pressure and 973/993 K temperature [17].