HRSG Systems and Implications for CCGT Plant Cycling

OMMI (Vol. 2, Issue 1)

April 2003

Background to the Design of HRSG Systems and Implications for CCGT Plant Cycling Fred Starr, European Technology Development , Surrey, UK Mr Fred Starr works as Principal Materials and Energy Consultant with European Technology Development. He is currently working on projects dealing with the cycling of power plant, materials for waste incineration, biomass combustion, and the development of high efficiency micro cogen systems. He previously worked at ERA Technology on high temperature corrosion and is a recognised expert in this field. Prior to this much of his earlier career was spent at British Gas working on advanced energy conversion systems including coal gasification processes and indirect fired gas turbine plant. [email protected] Abstract The principles which underlie the design of HRSG boilers for CCGT plant are outlined together with their effect on the susceptibility of HRSGs to plant cycling. The interaction between gas turbine and HRSG thermodynamics are described. Despite the marked increase in gas turbine inlet temperatures that has occurred over the past twenty years, there has only been a very slow rise in flue gas temperatures into HRSGs. The consequences of low flue gas temperatures is that HRSG plant is large in terms of steam generating ability and have also resulted in the need to generate the steam from two or more separate evaporators at different pressures. Boiler pinch point temperatures need to be in the 4°-12°C range for efficiency reasons, with most of the tubing in the HRSGs being finned for this reason It would seem that the preoccupation with heat transfer and HRSG size reduction has caused some additional problems. These are now well understood by designers and operators. However even during normal start ups HRSGs will be susceptible to thermal fatigue and flow assisted corrosion. In older units where condensate drainage was poor, or where restarts after plant trips were excessively quick, severe thermal fatigue occurred due to blocking of tubes and condensate quenching of headers. Although the increase in HRSG steam temperatures has been fairly slow, they are now reached the point where P91 steels are being used in the superheater sections. The benefits that this class of alloys confers and its potential shortcoming are briefly described. There are some other materials, which are coming onto the market, which may be beneficial in construction. 1.

Introduction

Since 1976 no coal fired plant have been built in the UK. Every power plant has been of the CCGT (Combined Cycle Gas Turbine) type using natural gas as a fuel. To simplify, a CCGT consists of a gas turbine, which produces about two thirds of the power of the plant. The exhaust from the gas turbine is used to produce superheated steam, in a HRSG or Heat Recovery Steam Generator, which then feeds into a steam turbine set, to produce additional electrical power from the plant. On modern units the steam and gas turbine are linked on the same shaft. Older designs might have one big steam turbine running off the steam from a number of HRSGs, which were in turn fired from a number of small gas turbines.

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When older power engineers get to hear about the HRSG part of the CCGT system, I am sure many of them have tended to dismiss this part of the plant as being just another steam boiler, and concentrated on the more exciting part of the technology; that is the gas turbine. Indeed, apart from specialised meetings on HRSGs, such as the one being organised by ETD Ltd in November of the this year (2003), the issues connected with HRSG materials and construction tend to be neglected by the mainstream conferences on boiler plant and steam turbine design, particularly in Europe. In consequence the idea for this background article emerged during two recent surveys of plant cycling by ETD Ltd, one on steam plant and the other on CCGT plant. This paper represents a personal review of HRSG design, garnered as much from person contacts as from the literature, and for that reason, a bibliography has been included, rather than sets of references, although where relevant some individuals are mentioned by name.

Exhaust from Gas Turbine

1 Inlet Duct 2 Distribution grid 3 HP Superheater 1 4 Burner 5 Split Superheater 6 HP Superheater 2

7 CO Catalyst 8 HP Steam Drum 9 Top Supports 10 SCR Catalyst 11 LP Steam Drum 12 HRSG Casing

13 Deaerator 14 Stack 15 Preheater 16 DA Evaporator 17 HP/IP Economizer

18 IP Evaporator 19 IP Superheater 20 HP Economizer 21 Ammonia Injection Grid 22 HP Evaporator

Figure 1: Schematic of horizontal HRSG boiler and superheater showing individual components (Courtesy Nooter Eriksen) Figure 1 shows a typical configuration for a horizontal HRSG. There are obvious differences with the pulverised fuel boiler shown in Figure 2. The HRSG does not have a furnace for steam generation, of course, all the heat coming from the gas turbine exhaust. Whereas, in a steam boiler, the superheater is located after the evaporator, in an HRSG the evaporator(s) are located downstream of the superheater. The inset table shows that there are actually three sets of evaporators, an HP set (22), an IP set (18) and a deaerator evaporator (16). There are also separate sets of HP and IP economisers. In a steam plant, although there are HP, IP and LP turbines, all the evaporation takes place at just one very high pressure. These differences can be traced back to heat transfer considerations, which have tended to dominate the thinking of HRSG designers. The HRSG has, therefore, required a completely different approach to design. It might be argued that in a coal or oil fired steam plant; heat transfer can look after itself, as temperatures in the furnace, where evaporation takes place are around 2000°C, and in the superheater and reheater region temperatures run between 800° and 1400°C. Indeed there is so much heat in the flue gases, after they leave the economiser, that this must be recovered in an air preheater! A big design issue in steam plant is the materials schedule, because of the risk of creep and fireside corrosion. Under specification will result in premature failures of water wall and superheater tubing.

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In an HRSG, it has been flue gas side heat transfer rather than materials, which has governed the design. A pessimistic view of heat transfer rates could result in a unit, which is too expensive, as even well designed HRSGs tend to be big in proportion to their output, because of the poor heat transfer rates and the need to minimise pressure drops along the duct. The positioning and dimensioning of heat exchanger surfaces is critical, as mistakes could lead to an HRSG working extremely badly. This need to concentrate on heat transfer has meant, some would argue, that plant operability and the consequential materials problems have not been given the prominence they deserve. The purpose of this article is to indicate the governing factors in HRSG design, and how these have affected failure mechanisms in HRSG systems, particularly in plant, which is subject to two shift operation, that is the starting up and shutting down of a plant on a daily basis.

Fig 2: Typical Pulverised Fuel Boiler from Mitsui Babcock Ltd

2. CCGT Gas Turbine Design Considerations Gas turbines intended for CCGT plant, run at significantly lower pressure ratios than those used in the aerospace sector. This has a number of advantages, which more than compensate for the one major shortcoming of this approach, which is that efficiency is 5-7 percentage points lower than it might be. A lower pressure ratio simplifies the design of the compressor, which is particularly critical since an industrial machine needs to rotate at the same rpm, no matter what the power output. Output from the gas turbine, that is specific power (i.e. power divided by airflow), is higher than if the machine is working at optimum efficiency. Although somewhat abstruse thermodynamic

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Turbine Inlet and Exhaust Temperatures °C

considerations come into play, a relatively low pressure ratio gives a low inlet temperature into the combustion turbine, so that more fuel can be burnt. Figure 3, based on some simple calculations, shows how the effect of increased pressure ratio tends to hold down the turbine exhaust temperature, so that above a pressure ratio of 15 the rise in outlet temperature is very slow. In fact if pressure ratios were fixed outlet temperature would rise in direct proportion to the inlet temperature, and gas turbines working at high temperature would be very inefficient.

1600 1400

Exhaust Temperatures

1200 1000

Inlet Temperatures

800 600 400 5

10

15

20

25

30

Pressure Ratio

Fig 3: Combustion Turbine Inlet and Exhaust Temperature versus Pressure Ratio

This approach also benefits heat transfer into the HRSG, which is yet another reason for the differences between industrial and aerospace machines. A low pressure ratio will result in a fairly high outlet temperature in to the HRSG, which become a vital factor in achieving reasonable rates of heat transfer when heat transfer coefficients are low. Low pressure ratios are advantageous in other ways. HRSG size is held down since, because of the high specific work, the volume of exhaust flow is reduced. HRSG efficiency is also improved due to the reduction in the levels of excess air going into the gas turbine. Even so, the excess air is about 3 to 4 times that needed for complete combustion. This mass of excess air, when it leaves the HRSG, carries with it a sizeable amount of heat energy, even though a typical stack gas temperature is only 100°C. In effect, the stack or chimney of a CCGT plant, which is of an excessive diameter when compared to that of a steam plant, represents inefficiency in just the same way as does a cooling tower. Any reduction in excess air will therefore improve efficiencies. In practice the designers of gas turbine systems have quite a lot of latitude in deciding how much power to take out of the gas turbine, so that as Figure 4, taken from manufacturer’s data, shows exhaust temperatures have tended to group within the 500-600° C band, with just a slight indication of an upward trend with pressure ratio. The two red points refer to aerospace derivative gas turbines, which because of the extremely high-pressure ratios have quite low exhaust temperatures. Such units are best used for standby power generation, the driving of pipeline compressors, or in niche installations where there is a requirement for relatively low grade steam. The sky blue square, indicating an exhaust temperature of 640°C at a pressure ratio of 30, is showing, perhaps, the shape of things to come, and refers to the GT24/26 ABB/Alstom reheat design of gas turbine.

Exhaust Temperature°C

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700 650 600 550 500 450 400 5

10

15

20

25

30

35

40

Pressure ratio

Fig 4: Actual Combustion Turbine Exhaust Temperatures v. Pressure Ratio

3. HRSG Geometries The HRSG evaporators, boilers, superheaters and, where they are fitted, reheaters are located in a very large rectangular duct. To reduce the velocity of the gas turbine exhaust, the duct opens out in a vee shaped manner, in both the horizontal and vertical directions, after the gas turbine, otherwise the pressure drops through the HRSG would be excessive. Swirl vanes are located after the turbine, as the flue gas will tend to corkscrew up the duct with some force, particularly when the gas turbine is working off design. The HRSG section of the ductwork can be laid out parallel to the ground, making the HRSG a horizontal type model. This type of duct leads into a stack, which will be equipped with silencers as was shown in Figure 1. The alternative vertical form of HRSG incorporates the boilers and superheaters within the stack structure. It follows that in a typical horizontal HRSG the heat transfer tubing is transposed in vertical harp type arrays across the duct. In the vertical form of HRSG the tubing can be in the form of loops which cross the duct horizontally as in a steam plant tower boiler. Where harps are used, these too will cross the duct in a basically horizontal manner. It will be clear that these two arrangements will drain and fill quite differently. The ducting for HRSGs is big because of the huge amounts of air and combustion products going through the gas turbine , and is therefore a design and maintenance issue in its own right. Mass flows for big gas turbines are over 600kg/s, but to put things in more concrete terms, just one of the big industrial gas turbines has more than enough power to propel a 747 type Jumbo Jet. Horizontal HRSGs can be up to 25 metres high and are up to 60 metres long. Thermal expansion of the ductwork is a significant design issue because of the size and the need to preserve the integrity of the internal insulation. With vertical HRSGs the flue gas has to turn through a right angle up into the stack soon after it has left the gas turbine. Very little heat will have been lost from the flue gas at this point and it is an area where thermal expansion problems can become quite difficult.

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4. Steam Generation in HRSGs 4.1 Duel or Triple Pressure Steam Generation It is a truism for both steam and gas turbines, that to attain the highest practicable efficiencies, the turbine entry temperature needs to be as high as possible. Steam temperatures of most HRSG units lie in the 480°C to 530°C range. To make use of a steam temperature of the order of 500°C or more, the steam must be expanded through an adequate pressure ratio. Since the pressure into the condensers is more or less fixed, this requires an adequate steam pressure into the turbines. For an HRSG equipped with a superheater and no reheat system, it can be shown that for a steam temperature of 500°C, the steam pressure would need to be around 60-80 bar. For one of the newer plants, equipped with reheat, pressures would need to be more than 120 bar. Generating steam at these pressures has a huge impact on HRSG designs, basically because much of the heat uptake happens when the water actually boils. In essence we are dealing with latent heat effects. Boiling point changes with pressure, and for water at 60 bar pressure, boiling occurs at 275°C. At 100 bar pressure, the boiling point is 311°C, and at these subcritical pressures, the water side temperature does not change until all the water has boiled into steam. We now start to see one of the problems with steam generation in HRSGs. Given that the flue gas into an HRSG is around 600°C, and the bulk of the heat uptake occurs at around the boiling point, this would mean that the temperature of the flue gas going up the stack would be just over 300°C. In other words, about half of the available heat in the flue gas would be lost up the stack, in a HRSG that produced steam at just one very high pressure, as in a pulverised fuel boiler.

HP Boiler and Superheater

LP Boiler and Superheater To Stack

From Gas Turbine

HP Turbine

LP Turbine

Alternator

Figure 5: Schematic of HP and LP Steam Production and Flows in a CCGT Plant

This problem is overcome by installing a further set of evaporators, economisers and superheaters in the HRSG, down stream of the high pressure system. This additional set produces steam at a much lower pressure, somewhere between 4 and 10 bar, the actual value being that which corresponds to the exit pressure from the HP turbine, or if the plant is fitted with a reheater, the IP turbine. As the

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boiling point of water at this sort of pressure is in the range 140°-180° C low temperature heat in the flue gases can be picked up quite easily. In addition some superheating of the LP steam is done with the aim of matching the temperature of the “cold” steam from the HP or IP turbine. It is then possible to merge the steam from these two different systems and put them to the LP turbine as shown in Figure 5. Steam, for deaeration, at an even lower pressure, can also be raised using the last vestige of heat in the flue gases. To summarise in a modern HRSG steam would be raised at three different pressures with separate evaporator and pumping circuits for each. This is in complete contrast to a pulverised fuel steam plant, where all the evaporation takes place at one very high pressure. Another difference is that feedheaters, of the type used on conventional pulverised fuel plant, using steam extracted from the turbines are not used on HRSG systems. This would simply raise the water temperature at the inlet to the economisers, and would reduce the amount of heat, which could be transferred from the flue gases. 4.2 Pinch and Approach Point Effects In extracting heat from the flue gas, the ideal is that the flue gas temperature should drop at a constant rate all the way down the duct. This would be simple to arrange if one was trying to heat another gas, or a liquid which did not evaporate as it temperature increased. It would then be easy to keep a constant temperature difference between the flue gas and the fluid being heated at every point. This is not too difficult in the economiser or superheater. In the former, the water stays as a liquid. In the latter, it stays as steam. It is impossible to maintain a constant temperature difference in the evaporator. Here as the water passes the tubing, the water turns to steam, with the temperature staying the same all the way through. There is no great problem with heat transfer at the exit to the evaporator. Assuming that the evaporator in question is at 100 bar pressure, the bulk water temperature at the exit will be the value stated previously, 311°C, with the tube wall temperate being just above this. The flue gas temperature at the entrance might be 450°C, so that the temperature differential would be around 70°C. As the flue gas moves through the evaporator section, it drops steadily in temperature, although the tube wall temperature in the evaporator would be just above 311°C all the way through. In theory, if the evaporator unit was infinite in size, there would come a point where the flue gas and water temperatures would be equal. This would give a “pinch point” of 0°C. In practice to keep the size of the evaporator realistic, the designer will adopt a pinch point of somewhere between 4°and 8°C. This is far lower than can be achieved in an HRSG on a refinery and it is one reason for the size of HRSG systems. The flue gas leaving the evaporator is likely to enter the economiser. If the same temperature differential between the flue gas and the water were maintained, as in the evaporator, boiling would tend to occur at the exit to the economiser. This would result in a “steaming economiser” with the possibility of the flow being blocked. It will be apparent that if the water flow were to be become blocked, all the water in the economiser would begin to turn into steam. Tubes would be subject to a mild degree of overheating but the main problems would be water hammer and tube-to-tube differential expansion. Accordingly, the outlet water temperature in the economiser is kept several degrees below the saturation temperature. This can be done in several ways, for example by cutting down the economiser heating surfaces and interposing part of the LP superheater system between the HP evaporator and economiser. The effect of this is to give an “approach temperature” between the flue gas and the evaporator outlet water temperature. The approach temperature might be 4°C, so this when added to a pinch point temperature of 8°C, would give an overall temperature differential of 12°C, greatly reducing the risks of boiling. See Fig 6.

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600 Flue Gas

Pinch and Approach Point Temperatures

500 400

Superheater

300

Evaporator Economiser

200 °C

Distance along HRSG

Figure 6: Schematic showing pinch and approach points for the HP superheater, evaporator and economiser tube banks As might be expected with such tight pinch points a slight change can have massive effect on boiler costs. Van Manen indicates that in going from an 8°C to a 6°C pinch point, the heat transfer surface would need to increase by about 10%. However the effect on steam turbine output, which is only about a third of the overall power output of the plant, is quite marginal, giving an increase in power output of 1% or about 1MW for the case described. This may not be the complete story, as extracting more heat in the HP evaporator would reduce the amount of heat downstream of this unit. For this reason, that is heat taken out by a tube bank located in the high temperature section will result in less heat being available in a lower temperature section, an HP economiser needs to be split into two or more sections. These cooler sections would be located much further along the HRSG, after the IP and LP evaporators. In this way a good amount of heat is available for heating of IP and LP steam, with lower grade heat being left for low temperature sections of the economisers. It will also be apparent that all the water in the system has to be heated, from near ambient temperature, after it has been returned from the condenser, with separate boiler pumps feeding each part of the system.

Superheater

Economiser

Fig 7: Effects of thermal expansion mismatch on siamised header arrangement for superheater and economiser

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There will therefore be some conflict in the location of the HP, IP and LP economisers and LP and IP superheaters. To simplify the construction, a siamised or common header may be used, with one half of a cross duct exchanger being made up of an economiser and the other half, the superheater. During startup the superheater tubes are likely to heat up faster, as they will be free from steam at this point in time, whereas the economiser will always be full of water. The superheater tubes will be put under a compressive load; the economiser tubes will be strained in tension, as shown schematically in Figure 7. Figure 8 shows a before and after picture of what happened after the bottom header was cut through at the point where the two headers had been joined

Fig 8: Photograph from Tetra Engineering Inc showing height difference After two crosses duct tube banks were separated

4.3 Heat Transfer and Tube Finning The temperature differential between the flue gas, steam and water, especially around the back end of the evaporators, is very poor. This situation is compounded by the poor heat transfer coefficient on the flue gas side. Water side and steam side coefficients are much better as will be seen in Table 1 Table 1: Typical HRSG Heat Transfer Coefficients Section of HRSG Heat Transfer Coefficient (W.m-2.K-1)

Flue Gas

Water in Economiser

Water in Evaporator

HP Steam

50

500

2500-10000

1000

It follows from this that tube wall temperatures tend to run quite close to the water and steam side temperatures. Even when temperature differentials are at their highest, the heat transfer rates are very modest. For example with a 100°C difference between the flue gas, as at the beginning of an HP evaporator, the actual heat transfer rate will be not much more than 5-10 kW.m-2. This compares with peak heat transfer rates in pulverised boilers which can be well over 200 kW. m-2. Given such poor flue gas side heat transfer rates, tubes must be of small diameter, with tight spacings and be of the finned type to provide sufficient heat transfer area. Figure 9 shows a set of photographs taken during the refurbishment of an evaporator by EH.Wachs Inc, which illustrates

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just how tight everything is. The only section of the HRSG which might not use finned tubes is the HP superheater where there might be a possibility of oxidation of the finning.

Figure 9: The tightness of construction of the tubing and finning in an HRSG evaporator is shown in during this header replacement by operatives from Wachs Inc. High efficiency finning is desirable as this reduces the size of the HRSG. Fin material with a high conductivity is needed. One manufacturer, Innovative Steam Turbines Inc, makes a unique form of once through HRSG, which uses Incoloy super-stainless tubing throughout, but opts for carbon steel finning wherever possible, as this material has a higher conductivity than that of the tube material. Fin shape and pitch are also critical due to the need to prevent excessive pressure drops through the system, otherwise gas turbine output will suffer. The overall pressure drop across the whole HRSG should not be much more than 25 mbar. A potential problem with finning, where different types of materials are employed, is expansion differences leading to thermal fatigue. There are also problems in welding fins to materials which tend to form bainite and martensite after cooling from weld temperatures. 4.4 Heat Transfer and Corrosion in Drum and Once Through Boilers As pressures and temperatures have become higher, the once through boiler has become more attractive. Once through designs avoid the need for a steam drum to separate the steam and water mixture after it leaves the evaporator. High pressure drums require very thick walled sections, which in a plant that is two shifting will be susceptible to thermal fatigue. This is probably a more vital factor in CCGT than in conventional steam plant. When a gas turbine “trips” and is then brought back on line, the HRSG is subject to a blast of cold air which causes condensation in the superheaters and causes steam production to halt extremely quickly. In a steam plant, trips of this type are unlikely as the furnace has a large amount of residual heat.

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Log Heat Transfer Rate

Mode of Boiling

Water Only

Water plus Steam Bubbles

Steam

Fig 10 : Modes of Flow in Water and Steam- Water Mixtures and Effect on Heat Transfer

The once through boiler is sometime termed the “Benson” type, wrongly it would seem if one goes back to Benson’s 1930’s conception. His ideas were based on the generation of steam under supercritical conditions, that is at pressures above 221 bar. Most once through designs, particularly for CCGT plant are subcritical. Here, as the temperature changes, the water in the evaporator section changes from water, in which steam bubbles are generated at the wall surface, through to steam going through several different flow geometries in the process. See Fig 10. During base load operation the steam exiting the evaporator section will have a mild degree of superheat and should enter the superheater proper in a completely dry condition. During start up, the steam from the evaporator will not only be saturated but will also contain free water. This needs to be removed in a separator. It will be seen that the changes in the modes of flow affect the heat transfer rate, the highest rate being when steam bubbles are tending to nucleate throughout the tube cross section. This is termed nucleate boiling. But at some point what is effectively a column of water with some steam must give way to a tube full of steam, which is on its way to becoming superheated. The heat transfer designer needs to ensure that steam blanketing does not occur, as shown in the last figure, since heat transfer rate will drop precipitously and tube temperature will rise. This is termed “departure from nucleate boiling”(DNB). Fortunately the extreme tube overheating that does occur in steam plant evaporators, and which actually promotes this “dry out condition” is unlikely in HRSGs, as tube wall temperature and external heat transfer rates can never reach really dangerous levels. However, during start up, the flow in the IP and LP economisers and evaporators can be of the two phase water/steam type, in both once through and drum types of HRSG. This results from the fact that since little HP steam has been generated, the flue gases reaching the back end of the HRSG will still be quite hot. There is then some potential for damage by water hammer and the development of tube-to-tube temperature differences. Water hammer is an obvious phenomenon. It can be heard! Once through boilers are subject to a more insidious form of damage called Flow Assisted Corrosion (FAC), since they have to use an All Volatile Treatment (AVT), to avoid deposition of solids in the evaporator. Unfortunately the use of AVT tends to lead to the dissolution of protective magnetite. Tube surfaces are then susceptible

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to erosion by two phase, high velocity flows, leading to FAC, as can occur in the low pressure units of the HRSG. Oddly enough the HP sections of an HRSG are unlikely to be affected by FAC since the difference in specific volume between steam and water falls rapidly at high pressures. It is this concern about steaming and the consequential FAC, which has meant that some designers prefer to stay with a steam drum on the LP units.

5. HRSG Design and Construction and Implications of CCGT Cycling 5.1 Plant Damage due to Normal Thermal Cycling Even during normal start ups, in which steam pressures and flow are allowed to develop in a reasonably well controlled manner, the tube banks in an HRSG are subject to temperature cycling and thermal fatigue. For example, unless steam from a package boiler is available, there will be no flow through the superheaters during the early part of start up. Temperature differentials will also develop across those tube banks containing water. Headers will run cooler than the tubing until flows start to develop. In principle, the heat transfer engineer can estimate the magnitude of these effects, which then can be used by mechanical engineers to estimate the fatigue life of equipment. As might be expected, rapid startups, caused by bringing the gas turbine on too quickly, will be detrimental to HRSG life. As the equipment heats up, the damage caused by an over rapid start will increase, as creep fatigue as well as fatigue becomes an additional life threatening mechanism. HRSG manufacturers have begun to supply ramp rate curves for HRSGs showing the level of plant damage. However Pearson has pointed out very strongly that curves, based on starting up from cold, do not tell the whole story, and this issue is addressed in the next section. 5. 2 Damage due to Condensate Quenching The accent on heat transfer has led to HRSG designers being preoccupied with the size and cost of the units. Naturally there have been commercial pressures too, but all good engineers will try to identify what are principal features of a technology that add to cost. Until quite recently steam temperatures and pressures were such that even the HP superheaters could be fabricated using T22 (2.25Cr-1Mo) steels, although header section thicknesses were becoming excessive. Partial penetration welding was also a cost saving option which could be highly advantageous. Tube spacing, into headers was and is such that full penetration welds were difficult. Design decisions such as this will reduce fatigue life even during well-managed start-ups. One particular batch of start up problems stemmed from the use of small-bore bottom headers in the superheater banks. The problem was compounded by the use of badly sized and located drain points, as large sized drain points start to become a stress issue as header diameter falls. The basic cause seems to have originated at an early design stage, where the view seems to have been taken that the headers should be made with as small a diameter as practicable, as the steam or water would flow straight from one tube to the other. See Fig 11.

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Fig 11: Flow situation in bottom header. Note that drainage space is highly restricted

This is fine during base load operation, but during plant trips and subsequent purges of the HRSG duct to remove inflammable gases, condensation of steam will occur in the HP superheater. The condensate must be drained off before the plant can be properly restarted. Narrow diameter bottom headers of the type shown are essential unsuitable for this task. Figures 12 and 13 shows the kind of build up that can occur, in one case with the drainage point in the centre, in the other with the point at one end. Figure 14 shows that if the drain point is too small, the condensate will back up uniformly to a near constant depth across the header.

Figure 12: Small diameter header with central drain point

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Figure 13: Small diameter header with end type drain point

Figure 14: Big header with small drain point

If the condensate is not fully drained off, the subsequent performance of the equipment, during restart, becomes an interesting exercise in heat transfer and damage to plant! The upper part of those tubes that are blocked with condensate will tend to overheat, as there is no steam flow. Tube to tube temperature differences will develop leading to stressing at the tube-to-antler connections. If the welding of these is of the partial penetration type the life of these could be very short.

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Eventually the condensate will start to boil and, due to the generation of steam, relatively cold condensate can be carried over as a mass, quenching the inside of the header. Heat transfer rates in such situations have been estimated at 8000 kW.m-2. Interligament cracking is a real possibility, and again the close spacing of the connections and use of headers with a heavy wall thickness will not help. Whether or not carry over of condensate is a problem can be ascertained from the changes in steam temperature during start up. If the increase in temperature is erratic, with sudden drops occurring, the chances are that condensate is being carried over. A longer time must be allowed to drain off condensate, and the ramp rate of the gas turbine should be reduced.

6. Current and Future Developments with HRSG Plants The exhaust temperatures of gas turbines are now over 600°C, and this should enable heat transfer specialists to take a more relaxed view about the design of HRSGs. Steam temperatures are very much higher than in the early eighties, now up 565°C. At these exhaust temperatures there is sufficient spare heat in the flue gases that it is worth reheating the steam. It follows that the output of the steam turbines are higher and HRSG heat rate is down. There is then perhaps a little less need to struggle for what is the equivalent, in steam plant terms, of a high boiler efficiency, which can help to offset any deficiencies caused by poor steam conditions. More thought can be given to part load and load following. HRSG pressures and temperatures now mean that T22 grade material (2.25Cr-1Mo) is no longer adequate for HP superheaters and headers. As in steam plant, designers are now turning to the T91, for the tubing and P91 for headers and connecting pipework, because of their high yield and creep strength at temperature. In some respects these materials, being 9Cr-1Mo martensitic steels, are better suited to HRSGs than to conventional steam plant. The reduced weight of tubing and headers should reduce start up times. The comparatively low alloy content of T91 is not an issue when ash type fire side corrosion is not present. The flexibility given by a reduced wall section is more beneficial in HRSG tube banks, which tend to be built out of straight length of tubes, compared to the type of construction in steam plant, where the superheaters are of the looped tube or pendant type. But the move upwards in temperatures will bring additional risks and P91 is not without its shortcomings. Operators will to have exercise even more care during start up and restarts, because high exhaust temperatures could be quite damaging. It would appear that the corrosion resistance in steam is very composition dependant The rate of steam side oxidation does not seem to be too damaging below 600°C, but here again it is easy to be misled by base load heat transfer calculations, which would indicate for a temperature of 565°C, the tube wall temperature would not be, in all probability, much above 580°C. During start up if the gas turbine is brought on line too quickly, temperatures could well exceed 600°C, albeit for short periods, since steam flows may not be well developed at this stage. An even bigger issue is that of oxide spalling due to temperature changes during cycling. This is probably more of an issue with HRSG superheater and reheater tubing than is steam plant because the very severe temperature changes which occur during plant trips. P91 is also likely to suffer from Type IV cracking. This is likely to be a serious issue in HRSG superheater banks, because as was mentioned, the tubing in many HRSGs is of the straight rather the looped type. There could also be problems in the welding on of fins onto P91, although it is not clear at this stage whether there any HRSG superheaters currently using finned P91 tubing. Where temperatures are not so high it may be worth considering the use of a modified 2.25Cr –1Mo steel called T23 which uses tungsten rather molybdenum and uses niobium and nitrogen to form strengthening precipitates. The welding of this is simpler than with T91.

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7. Conclusions The aim of this paper was to outline the thinking behind the design of HRSG systems and to explain some of the background to the engineering and materials problems that have been encountered. The paper has tried to show how the design has been dominated by heat transfer concerns, and that this has led to units that are difficult to build, inspect, and repair. The preoccupation with heat transfer has led to “clever” cost saving design solutions, such as the use of small headers and the neglect of system drainage. Also, one feels, heat transfer based calculations, may be leading to some over confidence in the promotion of ramp rate based life assessment programs. Fortunately the problems with condensate drainage, and the serious effects that this can have on plant, are now well recognised by both designers and plant operators. The gradual increase in gas turbine exhaust temperatures should give designers more latitude, so that it should be possible to build in more flexibility in load following capability. The paper has also touched on the background to flow assisted corrosion in HRSG boilers during start up, which again can be traced back to issues associated with heat transfer. FAC is a serious problem in air-cooled condenser plant, where the use of low chrome steel seems to be the solution. It may be that this could be an alternative to sophisticated water treatment approaches, which can be difficult to apply in plant that is two shifting. The move to higher temperatures and pressures will also imply the need for improved alloys, with P91 being the material of choice at present. This seems a much better option than the older grades of steel despite its known and potential shortcomings, but this is something which needs to be properly evaluated. Finally the author would like to say, that despite the under current of criticism on the over zealous application of heat transfer technology, as a long term member of HEXAG, a UK based heat transfer body, he has the greatest admiration for specialists in this field, especially those working with compact systems of the HRSG type. If anything, this paper was written with the aim of getting generating plant personnel and materials specialists to understand the hard work which has gone into making HRSGs amongst the most efficient recovers of energy that are operating in the world today.

Acknowledgements The views in this paper are those of the author’s and are not necessarily those of ETD Ltd. However the author would like to thank Dr Shibli of ETD Ltd for his encouragement in writing this paper, and Mr A. Fleming of ETD for his helpful comments. The author would also like to single out Mr M. Pearson of Pearson Associates and Dr D. Bogart of Innogy for their helpful discussions on HRSG mechanical design and issues connected with HRSG plant cycling.

Bibliography D G Wilson “The Design of High Efficiency Turbomachinery and Gas Turbines” MIT Publications 1984

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Anderson R A and Pearson M, ‘Reliability and durability from large HRSGs’, pp 21-46, CCGT Plant Components: Development and Reliability, Professional Engineering Publishing Ltd., 1999. D Bogart “ Dealing with the Rigours of Cyclic CCGT Operation: An Operators Perspective” pp 33-45 Modern Power Systems Oct 2001 A Van Manen “HRSG for Optimum Combined Cycle Performance” Paper 94-GT-278 1994 Int. Gas turbine and Aeroengine Congress and Exposition, The Hague, Netherlands, June1994 A Fleming, R V Maskell, L W Buchanan and T Wilson “Materials Developments for Supercritical Boilers and Pipework” pp 33-78 Materials for High Temperature Power Generation and Process Plant Applications, ed Strang Institute of Materials 2000 HEXAG (Heat Exchanger Action Group): This is a UK/Europe based organised comprising people and organisations involved in compact and two phase heat exchanger design and construction www.hexag.hw.ac.uk