Supercritical boiler technology for future market conditions

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H30-K0018-X-X-7600 Textseite, engl. 2001-09 D97 H:\BENSON\Homepage\Parsons_2003.doc of 13

Supercritical boiler technology for future market conditions

Joachim Franke and Rudolf Kral

Siemens Power Generation

Presented at Parsons Conference 2003

Oct. 2003



1 Introduction

The requirements for environmental protection and operating economy in future steam power plants make high efficiency levels and operating flexibility a matter of course not only in the EU but also in increasing measure around the world. Existing technologies have currently enabled fulfill-ment of these requirements by pulverized-coal-fired power plants and in part also by power plants with circulating fluidized bed (CFB) combustion systems.

Higher efficiencies can be achieved only along the path of higher steam temperatures and pres-sures.

2 State of the art

Power plants operating at supercritical pressure and high steam temperatures were already being constructed in the 1950s (Fig.1). The 1960s saw a series of supercritical plants in the U.S. (such as those equipped with the universal pressure boiler) and in the last twenty years supercritical plants were used exclusively in Germany and Japan. The latter were designed for sliding-pressure operation and thus also fulfill the requirements for high operating flexibility and high plant efficiencies at part load. (Fig.2).

To date, CFB power plants have been used especially for smaller power output levels, generally with drum boilers. Plants up to 350 MW are in the meantime already in operation and several plants equipped with Benson1 boilers have also been constructed. Supercriti-cal plants for ratings above 400 MW are planned.

Power plants operating at supercritical steam pressure have already demonstrated their opera-tional capabilities and high availability over decades. The transition to steam temperatures of 600°C and higher is now a further major development step, which decisively affects many aspects of the design of the power plant, especially of the boiler. Whether the transition to these high steam temperatures is economical also depends not only on the choice of main steam pressure, reheat pressure and feedwater temperature, but also on the range of fuel. 1 Benson is a registered trademark of Siemens AG

Plant

19 / 56513 / 53732 / 560Reheater 2 bar/°C

76 / 56582 / 565109 / 560Reheater 1 bar/°C

357 / 649321 / 621304 / 600Main Steam bar/°C

907306260Steam Flow t/h

32512585Electrical Output MW

195919571956Comission Date

Eddystone Nr. 1Philo Nr. 6Chemische Werke

Hüls

Figure 1: World’s first supercritical Power Plants



To date, the focus was on material de-velopment for the superheaters and the thick-walled components for high steam temperatures. However, investiga-tions indicate that the wall heating sur-faces can become the limiting compo-nents for further increases in steam parameters. One reason for this is the increasing fraction of superheater heat to be transferred with increasing steam parameters.

3 Effect on design

- Size of heat exchange surfaces

Higher steam temperatures automatically diminish the temperature differences between the flue gas and steam, with relatively large superheater and reheater heating surfaces as a consequence. As higher tube wall temperatures also mean an increased tendency to fouling, corresponding heating surface reserves must be provided.

Feedwater temperature has a large effect on the size of the heating surfaces in the cooler flue-gas path. Values of 290°C to 300°C or higher are necessary for high-efficiency plants. As on the one hand the flue-gas temperature downstream of the economizer is set in the design case at roughly 400°C – the temperature window for DeNOx – and on the other hand the water outlet temperature from the economizer is limited to avoid steaming, the upstream superheaters must absorb more heat with increasing feedwater temperature. At higher steam conditions, especially at increasing reheat pressures, the exhaust steam temperatures from the HP section of the turbine and thus the reheat inlet temperatures also increase. While these temperatures are still approx. 320°C at a de-sign main steam temperature of 540°C, they already increase to over 350°C in a 600°C main steam temperature design and even up to over 420°C in a 700°C design. This considerably de-creases the temperature difference to the flue gas, with the consequence of still larger heating surfaces in the reheaters.

Under consideration of a cost-effective heating surface design, feedwater temperatures should not exceed 300°C, and HP exhaust steam pressures should lie in the range of 60 bar.

Avedorevaerket 2BoxbergSkaerbaekvaerketLippendorfNordjyllandsvaerketAghios DimitriosSchkopauNeckar 2RostockHemwegMeri PoriStaudinger 5Fynsvaerket

Tachibanawan Tachibanawan 1Haramachi 2Matsuura 2Nanao OtaShinchiNoshiroHekinan 2Shin Miyazu

415 915 410

2 x 930 410 350

2 x480 340 550 660 550 550 430

1050700

10001000500

1000600700450

332285310285310242285285285261240285275

285275280275275275293275270

582 / 600 545 / 580

582 / 580 / 580554 / 583

582 / 580 / 580540 / 540545 / 562545 / 568545 / 562540 / 540540 / 540545 / 562540 / 540

605 / 613570 / 595604 / 602598 / 596570 / 595542 / 567542 / 567543 / 569541 / 569

2001200019991999199819981997199719941994199419921992

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Power PlantOutput [MW]

Design Pressure *) [bar]

Steam TemperatureBoiler Outlet [°C]

Year ofCommisioning

*) max. allow able working pressure at boiler outlet

Japan

Europe

Avedorevaerket 2BoxbergSkaerbaekvaerketLippendorfNordjyllandsvaerketAghios DimitriosSchkopauNeckar 2RostockHemwegMeri PoriStaudinger 5Fynsvaerket

Tachibanawan Tachibanawan 1Haramachi 2Matsuura 2Nanao OtaShinchiNoshiroHekinan 2Shin Miyazu

415 915 410

2 x 930 410 350

2 x480 340 550 660 550 550 430

1050700

10001000500

1000600700450

332285310285310242285285285261240285275

285275280275275275293275270

582 / 600 545 / 580

582 / 580 / 580554 / 583

582 / 580 / 580540 / 540545 / 562545 / 568545 / 562540 / 540540 / 540545 / 562540 / 540

605 / 613570 / 595604 / 602598 / 596570 / 595542 / 567542 / 567543 / 569541 / 569

2001200019991999199819981997199719941994199419921992

200120001998199719951994199319921991

Power PlantOutput [MW]

Design Pressure *) [bar]

Steam TemperatureBoiler Outlet [°C]

Year ofCommisioning

*) max. allow able working pressure at boiler outlet

Japan

Europe

Figure 2: Large Supercritical BENSON Boilers in Europe and Japan - References



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Q [%]

1 2 3

RH SuperheatingRH Superheating

Preheating andEvaporation


HP SuperheatingHP Superheating

540 / 560 °C250 bar600 kg/s

600 / 620 °C290 bar543 kg/s

700 / 720 °C350 bar476 kg/s

Steam conditionsT HP/RHp HPMHP

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39

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RH SuperheatingRH Superheating



HP SuperheatingHP Superheating

540 / 560 °C250 bar600 kg/s

600 / 620 °C290 bar543 kg/s

700 / 720 °C350 bar476 kg/s

Steam conditionsT HP/RHp HPMHP

- End of evaporation

The location of the separator determines the location of the end of the evaporator on startup and at low load in recirculation mode. Usually the separator is configured such that its temperature is slightly superheated at the lowest once-through load point. Design of the boiler for high steam temperatures and pressures leads to this being already the case in lower areas of the furnace walls instead of as from the outlet first pass or in the boiler roof. The reason for this is the increas-ing degree of superheat and correspondingly decreasing fraction of evaporation in the heat input to the HP section with increasing steam parameters. At a load of 40%, the degree of superheat in a 540°C boiler is approx. 27%, and this in-creases to 39%, for example, in a design for 700°C main steam temperature (Fig. 3 and Fig. 4). As the highly loaded heating surface area must lie up-stream of the separators for reasons of evaporator cool-ing and the separator thus cannot be moved arbitrarily toward the burn-ers, a signifi-cantly larger degree of superheat will result at the lowest once-through operating point (Fig.5). This considerably increases the downward step of the steam temperatures on the transition to re-circulation mode. In order to extensively prevent this temperature change, the transition from once-through to recirculation mode must be placed at a very low load point, requiring recirculation mode only for startup. Whereas for boilers with spiral wound tubing the minimum load in once through operation is in the range of 30% to 40%, an evaporator based on the "Benson Low Mass Flux"[1] design with vertical rifled tubes enables loads to below 20%.

Furnace Design and Size is given by Coal and Ash Quality

Figure 3: Heat Flow Distribution in Variable Pressure Operation at 40% Load



Furnace Design and Size is given by Coal and Ash Quality

FEGT = IDT - 50 K FEGT = Furnace exit gas temperature IDT = Initial deformation temperature of ash

Zones of Evaporation (at Part Load)

Full Load Steam Conditions 190 bar / 535 °C / 535 °C




∆h Evaporation at 40% Load (sliding pres-sure)

0 100 200 bar1000

1800

2200

2600

3000

kJ/kg

Corresponding full load steam pressure: 350 bar 290 bar 250 bar 190 bar

Figure 4: Increasing steam conditions lead to different evaporator designs



Figure 5: Water and Steam Temperatures in the h-p Diagram

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0 20 40 60 80 100 120 140 160 180 200 220 240 260 280 300 320 340 360 380 400p [bar]

h [k

J/kg]

T [°C]

700

750

800

EvaporatorInlet

Nose

Water Wall Outlet

Superheater Outlet

Economiser Inlet

100% Load

Roof

40% Load

EvaporatorInlet

Nose

Water Wall Outlet

Economiser Inlet

Roof

703 °C / 358 bar476 kg/s

603 °C / 300 bar543 kg/s

544 °C / 261 bar600 kg/s

Reheater

100% Load

540 °C 200 bar

Water walls

The water walls in boilers for subcritical steam conditions are generally configured as evaporators. At increasing steam temperatures and pressures, the fraction of evaporator heating surfaces de-creases, with the result that parts of the water walls must also be configured as superheaters, i.e. downstream of the separator. In the highly loaded furnace area, spiral-wound evaporator tubing is usually used with smooth tubes and high mass fluxes – approx. 2000 – 2500 kg/m³s. As spiral-wound furnace tubing of this type is not self-supporting, it is reinforced with support straps which are welded to the tube wall with support blocks.

High steam parameters also lead to higher material loading in the evaporator. The previously ex-isting design reserves are no longer available, with the result that a detailed stress analysis is re-quired for the design of the evaporator tubing in each case. As a result of the requisite large wall thicknesses, the design of highly loaded heating surface areas is in part no longer determined by the primary stresses due to internal pressure but rather by the secondary stresses due to re-



strained thermal expansion. The higher evaporator temperatures also result in increasing tem-perature differences between the tubes and support straps on startup and shutdown. This in turn leads to longer startup times, especially on cold start.

The "Benson Low Mass Flux" design developed by SIEMENS with de-sign mass fluxes of approx. 1000 kg/m²s and below and with vertical rifled evaporator tubes requires no additional support structure and thus also does not im-pair plant flexibility in spite of wall outlet tempera-tures of approx. 500°C and above.(Fig.6). In a design for main steam tem-peratures of 600°C and above, the creep strengths of the wall materials commonly used to date such as 13CrMo44 (T12) are no longer sufficient, necessitating the transition to new developments such as 7CrMoVTiB1010 (T24) or HCM2S (T23). This is already the case at steam pressures of 300 bar and above for lower design temperatures. Looking at primary stresses the creep strengths of these materials, which require no post-welding heat treatment, permit steam temperatures up to 530°C in the furnace walls depending on main steam pressure, but the corrosion resistance and secondary stresses limit these values down to 500°C. Main steam temperatures of 630°C at mod-erate steam pressures are thus achievable as regards the walls.

At higher steam temperatures, materials such as HCM12 or T92 are required which must be heat-treated after welding. In order to minimize the manufacturing expenditure in such a design, the erection welds on evaporator tubes must be reduced to the absolute minimum possible. This is currently feasible only with vertical tubing. The relatively complex welds in the corners for spiral-wound furnace tubing are eliminated and the individual wall segments are welded together only at the fins. Welding of tubes may become necessary only in the horizontal plane. Solutions are also available for this which minimize expenditure on heat treatment on erection. In all cases, it can be stated that the problems in the design of the water walls increase dispropor-tionately with increasing steam pressures. A reduction of main steam pressure from 350 bar to 250 bar reduces the efficiency of a 700°C plant by 0.7 percentage points but it also reduces the wall outlet temperature from 540°C to 500°C and makes a design with materials without post weld heat treatment possible. Main steam pressures far above 250 bar should therefore be avoided, also in plants with high steam temperatures.

Simple, cost-effective manufacture and assembly of water walls

Simpler maintenance e. g. for tube damage

No stresses due to thermal expansion because welded-on support straps are eliminated

Reduced auxiliary power consumption

Reduced slagging

Low mass flux design with natural circulation characteristic

Figure 6: Vertically-Tubed Furnace for BENSON Boilers Principle and Characteristics



- Evaporator/superheater dividing point

At high steam parameters the water walls can no longer be designed entirely as an evaporator. The transition from evaporator walls to superheater walls then lies above the furnace. This transi-tion must be designed so as to minimize the temperature differences between the evaporator and superheater sections of the walls which automatically result on water filling after shutdown, espe-cially on water filling after an emergency shut down. Values of up to 80 K represent no cause for concern. For higher values such as can occur at very high steam conditions as well as in large furnaces, a flexible connection, not necessarily welded gas-tight, should also be taken into consid-eration for this transition.

- Superheater heating surfaces

For steam temperatures up to approx. 550°C, all heating surfaces can be constructed of ferritic or martensitic materials, while at 600°C austenitic materials are necessary for the final superheater heating surfaces for both the HP section of the boiler as well as the reheater. In addition to the strength parameters, corrosion behavior on the flue-gas and oxidation behavior on the steam sides is especially determinative for material selection. Fig.7, Superheater materials for high tem-peratures, shows a selection of available materials. With regard to strength parameters, construc-tion of superheater heating surfaces for steam temperatures up to 650°C is currently already fea-sible with austenitic steel materials. The corrosion resistance of the available materials however reduces the design limits to about 630°C.

X3CrNiMoN1713 595

615

605

635

580

580

620

600

AC66

Esshete

Super 304H (FG)

645 620NF 709

630 630HR 3C

EN

VdTÜV

VdTÜV / ASMEMITI

ASME / MITI

MITI

under development

TP 347 H (FG)

Save 25

Alloy 617 A130

620

655

685

600

630

720

VdTÜV / BS

VdTÜV / ASMEMITI

under develop-ment / MITI

Approved byMaximum HP Steam Temperature limited by

Creep Rupture Strength* Corrosion

* 100 MPa at Steam Temperature +35K

Figure 7: Available Superheater Tube Materials



- Thick-walled components

In the first steam generators with very high steam temperatures, austenitic materials were used for the hot headers and connecting lines. However, the poor thermoelastic behavior – low thermal conductivity, high thermal expansion – render these materials unsuitable for boilers which are im-plemented in power plants with a large number of load changes and minimum startup times.

The development of chromium steels such as P91, P92 or E911 has enabled steam temperatures up to 620°C without the use of austenitic materials for thick-walled components. More recent de-velopments such as NF12 and Save 12 could extend the limits of implementation at moderate main steam pressures up to 650°C in the near future.

With regard to the thick-walled com-ponents, espe-cially for the main steam headers, it proves that the main steam pres-sures should more likely lie be-low 300 bar for optimum compo-nent utilization (Fig.8) [2].

- Effect on operation

Power plants which are designed for fast load changes and short and frequent starts must neces-sarily be operated in sliding-pressure mode. Only then does the material loading of the turbine re-main acceptable: in sliding-pressure operation– usually between full load and 40% load - the tem-perature curve in the turbine remains nearly constant over the entire load range. These advan-tages for the turbine contrast with disadvantages for the boiler. For example, the temperatures in the water walls decrease from full load to part load by approx. 100 K. Due to their magnitude and the ordinarily larger wall thicknesses at the elevated steam parameters, the temperature changes during start up and load variations place increased requirements on the design of the thick-walled components such as multiple parallel passes, but also on the design of the tube walls, such as vertical tubing, in order to achieve similar startup times and load change rates to those in plants with conventional steam parameters.

With increasing steam parameters, the degree of superheat at the outlet of the evaporator sections of the water walls at the lowest once-through load point also increases. A high degree of superheat

Main steam pressure upstream of turbine [bar]

540

360

320

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200

Main steam temperature upstream of turbine [°C]

560 580 600 620 640 660 680 700 720

Ni-basedmaterial

Austenitic

X 20

P 91

E 911/NF 616

NF 12

TP 347H FG

Alloy 617

Ferritic

Figure 8: Optimum Main Steam Conditions with given Main Steam Header Dimensions



leads to a temperature reduction at the evaporator end and in the superheaters in the transition to recirculation mode. The separators are therefore moved as far as possible toward the burner zone. Operating measures to reduce the degree of superheat are increased excess air, flue-gas recircu-lation and use of the uppermost burner levels. The higher the steam temperatures and pressures become, the more important is the lowest possible load point in once-through operation, so that the once-through/recirculation mode transition need be traversed only on startup.

The large degree of superheat in the separator at the lowest once-through operating point also results in changes in startup behavior at high steam parameters. On warm and hot startup in recir-culation mode, the achievable hot steam temperatures are below the values required by the tur-bine. The earliest possible transition to once-through operation is necessary in order to shorten startup time, as full main steam temperatures are also already possible at low load in this operating mode.

High feedwater temperatures can restrict the sliding-pressure range in plants with very high main steam pressures. In order to prevent the economizer from approaching the evaporation point at low load, the pressure must be already fixed below 50% load or still higher depending on the design.

Increasing steam parameters also decrease the design reserves of nearly all pressure part com-ponents, as, not least for reasons of cost, the decision for advanced materials is not made until the reserves of lower quality materials become insufficient. This also increases the requirements on control quality: temperature deviations from the design value, such as on load changes, must be kept to a minimum. The conventional cascade controller is no longer sufficient for superheat tem-perature control; concepts such as two-loop feedback control or observer features provide significantly better control quality.

Special attention must be given to feedwater control. Conventional sys-tems which employ only simple delay modules to account for the dynamic differences be-tween heat release by the fuel and heat absorption by the evaporator tubes usually lead to large tempera-ture fluctuations at the evaporator out-let on load changes. New con-trol concepts which account for effects such as those of changes in the evaporator inlet tempera-

Evaporator outlet temperature

0 250 500 750 1000 1250 1500

Previous feedwatercontrol concept

New feedwatercontrol conceptwith allowance for- inlet enthalpy- storage of thermal energy

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sTime

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Figure 9: Comparison of Feedwater Control Concepts Load reduction from 100% to 50%



ture or the thermal storage capacity of the tube wall in the form feed forward control (Fig.9) in-crease control quality decisively and thus minimize the use of more expensive, higher-quality mate-rials.

For high degrees of superheat at the lowest once-through load point, the transition from recircula-tion mode to once-through operation and back can no longer take place without delay due to the relatively large temperature change; the control must be adapted accordingly for a sliding transi-tion.

4 Other effects

Design of the tube walls in particular is impeded by the high steam temperatures and pressures. The design parameters should be selected as best as possible so as not to necessitate the use of materials for which heat treatment must be performed after welding. A significant aspect for this is selection of the fuel. Coals with low ash deformation temperatures require large furnaces, associ-ated with high heat input to the walls. A 100K lower ash deformation temperature leads in a com-parable boiler concept to a temperature increase at the wall outlet of about 25K. Because of this for the currently available wall ma-terials without post-welding heat treatment, the ash deformation tem-perature for a 600°C boiler may not be much lower than 1200° (Fig.10).

The implementa-tion of flue-gas recirculation – extraction of the flue gases if pos-sible upstream of the air heater in order to reduce the negative effect on exhaust-gas temperature– can shift the limits to higher steam pa-rameters.

Steam generators for power plants with high steam parameters and hence high plant efficiencies are consequently also designed for high boiler efficiencies. The lowest possible exhaust-gas tem-peratures – 115°C to 110°C can be achieved depending on the coal – and lower excess air are prerequisites for this. Both of these factors lead to an increased heat input to the evaporator and thus impede the design of the wall heating surfaces.

400

500

600

1100 1200 1300

Ash deformation temperature °C

Wall exit temperature °C

TFD = 540°CpFD = 250 bar

TFD = 600°CpFD = 300 bar

TFD = 700°C pFD = 350 bar

13CrMo447CrMoVTiB1010P92A617

Figure 10: Design Limits for Water Wall Materials



The high tube wall temperatures of the superheater heating surfaces as well as lower excess air and low-NOx firing systems increase the corrosion problem. For the selection of supeheater mate-rials the resistance to scale formation from the flue-gas atmosphere and steam is therefore just as important as creep resistance.

4.1 Special aspects for CFB

The advantages of CFB technology are uncontested for low-grade fuels or for fuels with widely fluctuating quality as well as for low exhaust-gas emissions without post-combustion control meas-ures. CFB plants up to capacities of 350 MWe are currently in operation. However, only once-through operation with high steam conditions render CFB technology serious competition for pul-verized-coal firing. A plant for approx. 460 MWe with steam parameters of 560°C/580°C and 265 bar was developed in an EU research program. The BENSON "Low Mass Flux" design was se-lected as the evaporator concept. It fulfills the requirements of a fluidized bed to a special degree: the tube orientation parallel to the flue gas/ash flow ensures low susceptibility to erosion, and tem-perature variations between the evaporator tubes are extensively prevented, as non-uniform heat inputs are evened out by the natural circulation flow characteristic of the low mass flux design. It also features an especially simple construction, as flow through all of the tubes in a single pass is parallel, thus eliminating the need for elaborate water/steam distribution.

The suitability of this evaporator system for sliding-pressure operation also fulfills all requirements for a power plant with regard to operating flexibility.

4.2 Combined-cycle plants

Heat-recovery steam generators downstream of gas turbines are usually designed as drum boil-ers. Increasing exhaust-gas temperatures downstream of gas turbines as well as the increasing requirements on flexibility of a combined-cycle plant with frequent starts also make the use of once-through systems interesting here. Elimination of the drum on the one hand increases operat-ing flexibility and on the other hand is a noticeable cost aspect. In the Cottam combined-cycle plant, a heat-recovery steam generator with a once-through evaporator based on the Benson "Low Mass Flux" design was constructed for the first time and runs successfully in commercial operation since Sept.1999. This evaporator concept is characterized by extremely low mass fluxes which still lie far below those of fired boilers.



5 Summary and outlook

Steam temperatures of 600°C to 620°C are currently possible as a result of efforts in materials development. However, not only are new materials necessary for higher temperature ranges, but further development was also necessary for the wall materials. On further temperature increases, previous design concepts can no longer be adopted without modifications. New designs are nec-essary for the evaporator in particular in order to give boilers for high-temperature plants similar flexibility to that of previous once-through boilers.

The Low Mass Flux Design provides an evaporator concept which meets the new requirements and which permits further development to higher steam parameters for pulverized-coal-fired boil-ers and for boilers with circulating fluidized bed firing as well as for heat-recovery steam genera-tors downstream of gas turbines.

A further increase in steam temperatures appears possible in the next years with continuous ma-terials development, but without using nickel based materials not more than 10K to 20K. From the current standpoint, the jump to 700°C will not take place until the next decade. However, from an economic perspective, the high steam temperatures will only be selected given correspondingly competitive materials prices and if, among other things, the appropriate main steam and reheat pressures are selected and the fuel ranges are limited.

References

[1] J. Franke and R. Kral Innovative Boiler Design to Reduce Capital Cost and Construction Time Power-Gen 2002

[2] J. Franke, R. Kral and E. Wittchow Steam Generators for the Next Generation of Power Plants VGB Power Tech 12/99

Supercritical boiler technology for future market conditions

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