JI THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS 345 E. 47 St., Now York, N.Y.10017 The Society shall not be responsible for statements or opinions advanced in papers or in discussion at meetings of the Society or of its Divisions or Sections, or printed in Its publications. Discussion is printed only if the paper is published In an ASME Journal. Released for general publication upon presentation. Full credit should be given to ASME, the Technical Division, and the author(s). Papers are available from ASME for nine months after the meeting. Printed in USA. Copyright 1984 by ASME 84-GT-160 DESIGN OF A HEAT RECOVERY STEAM GENERATOR David R. Logeais, P.E. Engineering Manager Deltak Corporation Minneapolis, Minnesota Associate Member ASME ABSTRACT A gas turbine in the size range of 20,000 hp (14.9 MW) was retrofitted with a heat recovery steam generator (HRSG). The HRSG produces high pressure superheated steam for use in a steam turbine. Supplementary firing is used to more than double the steam production over the unfired case. Because of many unusual constraints, an innovative design of the HRSG was formulated. These design constraints included: A wide range of operating conditions was to be accommodated. 2. Very limited space in the existing plant. 3. A desire to limit the field construction work necessary in order to provide a short turnaround time. This paper will discuss the design used to satisfy these conditions. INTRODUCTION Two identical heat recovery steam generators were installed in the Gulf Coast plant of a major chemical company. They were to replace existing boilers which were being scrapped. Each HRSG recovers the exhaust heat from a 20,000 HP Westinghouse W-191 gas turbine. The turbine is fueled with gas and is in compressor drive service. Design steam conditions from the HRSG are 150,000 lb/hr (68,039 kg/hr) at 625 psig (43.1 barg) superheated to 750 degree F (399 degree C). DESIGN SPECIFICATIONS Basic criteria for the design of the heat recovery steam generator were established by the owners specification. One criteria was the wide operating range over which the unit was to function. The nominal design case was based on 100% turbine power output on an 80 degree F (27 degree C) ambient day. Figure 1 illustrates the flow conditions for this fired base case. At these conditions, the HRSG was to produce 150,000 lb/hr (68,039 kg/hr) of steam at 625 psig (43.1 barg) and 750 degree F (399 degree C). There was not enough energy available available by cooling the exhaust gases to produce this quantity of steam. Therefore, supplementary firing was required to increase the exhaust gas temperature from the turbine. It was also specified that the steam temperature should not fall below 730 degree F (388 degree C) when operating without supplementary firing. In the unfired mode steam pressure was to be maintained at 625 psig (43.1 barg). It was required that the system be designed for dual fuel capability. The supplementary burner was to be designed for initial operation on gaseous fuel. The fuel consists of a mixture of hydrogen and methane. Nominal composition (by volume) is 60% Hydrogen and 40% Methane. The supplementary burner and the HRSG were to be designed for future use of distillate fuel oil. SITE CONSTRAINTS Another difficult condition placed on the design was the space constraint caused by the requirement that the equipment furnished fit in a limited space in an existing plant. The HRSG is a series of heat exchangers. In this case they consist of boiler, superheater and economizer. As with any heat exchanger, decreasing its physical size results in an increase in fluid velocities and associated pressure losses. In this case a maximum of 10 inches (254 mm) water column back pressure was to be imposed on the gas turbine. The problem became one of making the HRSG small enough to fit in the existing space without imposing excessive backpressure on the engine. Related to the space constraint was the users desire to minimize field work and turn around time required for the installation of the HRSG. It was specified that the HRSG be factory assembled as much as possible.
C 382 F (194 C) Economizer Feedwater 57,500 LB/HR, 240 F (71,441 KG/HR, 116 C) 467 F (242 C) 545 F (285 C) 00 81,500 LB/HR, 889 F r----- ( 36,968 KG/HR, 476 C) Superheated Steam 150,000 LB/HR,889 F (68,039 KG/HR,476 C) (527 C) Superheater Slowdown Saturated Water Boiler 7,500 LB/HR, 500 F O (3,402 KG/HR,260 C) Saturated Steam 150,000 LB/HR, 500 F (68,039 KG/HR, 260 C) 1130 F (610 C) Desuper heater 81,500 LB/HR, 631 F 1(36,968 KG/HR,333 C) Superheated Steam 150,000 LB/HR,750 F 68,500 LB/HR,889 F (68,039 KG/HR,476 C) (31,071 KG/HR,476 C) Supplementary Fuel 87.56 MBTU/HR (25.66 Mw) Turbine Exhaust Gas 248.6 LB /SEC,770 F (112.8 KG/SEC, 410 C) FIG. I FIRED BASE CASE FLOW DIAGRAM Supplementary Burner DESIGN SOLUTIONS A natural circulation, finned tube evaporator was chosen. The evaporator was designed with inclined tubes. The superheater and economizer have finned horizontal tubes. As can be seen in Figure 2, the boiler, superheater and economizer are stacked one above the other with the exhaust gases flowing vertically upwards. The supplemental burner is mounted in a horizontal leg of the inlet ductwork with the exhaust gases turning 90 degrees before entering the superheater. Table 1 gives specific mechanical data on each heat exchanger section. The gas flow and equipment arrangement was chosen for a number of reasons. By stacking the boiler equipment and having vertical gas flow, space to the north side of the HRSG was conserved. The area north of the HRSG is occupied by a road which could not be blocked or rerouted. Having the equipment elevated provided maximum room for the supplementary burner and its associated ductwork. The natural circulation boiler with its inclined tubes was chosen for its advantages over a forced circulation unit that typically has horizontal tubes. These advantages include: 1. A higher circulation ratio. Forced circulation boilers typically have a circulation ratio between 3:1 and 7:1. A natural circulation boiler is designed for a circulation ratio of 15:1 or greater. The higher circulation ratio and inclined tubes minimize the potential for steam blanketing and resulting tube failure. 2. The natural circulation boiler is able to respond to changing load almost instantaneously. As the heat flux on any given boiler tube changes, the circulation is self adjusting to accommodate this change. For example, if heat flux (load) increases, more steam is generated, which in turn causes more vigorous circulation. The natural circulation boiler doesn't require a circulating pump. This simplifies the system and avoids the power requirement of the pump. Finned tubes were chosen for all heat exchanger sections. Finned tubes provide extended heating surface and can decrease the size of the HRSG by as much as 75% when compared with a unit using bare tubes. Finned tubing also permits lower temperature differences between the heated and heating fluids. The practical effect of this is to increase energy recovery by lowering stack temperatures. 2
II ECONOMIZER TUBES DUCT C H-H r-duct B 1-DUCT BURNER JO L (24.4M)', BOILER STEAM DRUM- {ILOWER DRUM W/ DESUPERHEATER^ neh DUCT A SUPERHEATER PERFORATED PLATE Lam- TURNING VANES SECTION A - A SUPERHEATER TUBES DUCT A V^ II i^ I^ H ^^I GRADE GENERAL ARRANGEMENT OF HEAT RECOVERY STEAM GENERATOR FIGURE 2 END ELEVATION SIDE ELEVATION
TABLE 1 - MECH ANICAL DATA Su perheater Boiler Economizer Heating Surface - ft 2 (m 2 ) Tube Diameter x minimum thickness - in (mm) 18,823 (1,749) 65,426 (6,078) 41,992 (3,901) 2x.135 (51x3.4) 2x.105 (51x2.7) 1-1/4x.105 (32x2.7) Tube material SA-213-T22 SA-178A SA-178A Number of tubes transverse to flow Tube CL spacing transverse to flow - in (mm) Number of tube rows Tube CL spacing in direction of flow - in (mm) Drum/Header diameter - in (mm) 22 88 32 4 (102) 4-1/4 (108) 3-1/4 (83) 4 16 10 6 (152) 4 (102) 3-1/2 (89; 8/10 (203/254) 42/30 (1067/762) 4 (102 Drum/Header material SA-335P11/SA106B SA-515-70 SA-106B Fins used in all sections are the segmented type. Fin height is.75 inches (19 mm), fin thickness is.05 inches (1.3 mm). The fins are spirally wound and continuously welded to the tube. Maximum fin density is 6 fins per inch (236 fins per meter). This fin density has been used successfully in gas turbine exhaust service where the turbine and supplementary burner are fueled by gas. It is judged to be suitable for the future addition of oil burning capabilities in the supplementary burner. As a precaution for this future condition, sootblower lanes were provided in all heat exchanger sections. Should they prove necessary, sootblowers can be added with minimal modifications. OPERATING CONDITIONS As previously mentioned, one design requirement was for the HRSG to operate over a wide range of conditions. This requirement had a major impact on the design and operation of the superheater and the economizer. As can be seen in Table 2, the final steam temperature from the superheater was to be a minimum of 730 degree F (388 degree C) in the unfired case. In this case the turbine exhaust gas temperature is 770 degree F (410 degree C). In the fired case the turbine exhaust gas must be fired to a temperature of 1130 degree F (610 degree C) to obtain the steam production desired. Yet under these conditions, the final steam temperature was only allowed to rise to 750 degree F (399 degree C). The solution was to design the superheater for the unfired case. By doing so it was considerably oversurfaced for the fired case. From Table 1 it can be seen that in the fired case the final steam temperature is 889 degree F (476 degree C). This temperature is then reduced downstream by the use of a desuperheater. Due to the high final steam temperature in the fired case, the superheater tubes had a design temperature of 1000 degree F (538 degree C). This high temperature necessitated the use of SA-213-T22 tube material. This is a specification for a seamless ferritic steel of nominal composition 2-1/4% Cr, 1% Mo. Another problem arises due to the wide operating range. The economizer is included to increase operating efficiency by providing a lower stack temperature. The economizer was designed for an approach temperature of 33 degree F (18 degree C) in the fired base case. Approach temperature is the difference between economizer outlet water temperature and the saturation temperature in the boiler. The lower the approach temperature, the more energy being recovered in the economizer. An approach temperature of zero degrees corresponds to the maximum theoretical energy recovery in the economizer. A typical problem with designing for an extremely low approach temperature is steam generation in the economizer at low loads. Steaming can cause water hammering, blockage of flow circuits and excessive pressure drop. In this case, the large turndown (more than 2 to 1 between the fired base case and the unfired alternate 2 case) could cause steaming in the economizer at low loads. The design solution was to provide an extra set of headers in the economizer. This is shown in Figure 3. In - -^ -- -- --T ^-- FEEDWATER INLET 40 % OF D HEATING SURFACE------... '.. t - VALVES HEAT ING --- SURFACE ( IIIIIID I ^^ _... - ---._ -- - - _.. _.----PTO BOILER EXHAUST GAS FLOW ECONOMIZER ARRANGEMENT FIGURE 3-4 -
II TABLE 2 - PERFORMANCE SUMMARY Operating Case II, f -i i Maximum Fired Power - % 100 100 80 Ambient temperature - F ( C) 80 (27) 80 (27) 40 (4) Supplementary firing - MBTU/hr (MW) 87.56 (25.66) 0 88.6 (25.97) Exhaust gas flow - lb/sec (kg/sec) 248.6 (112.8) 248.6 (112.8) 271.9 (123.3) Exhaust gas temperature - F ( C) 770 (410) 770 (410) 615 (324) Exhaust gas temperature after supplementary firing - F ( C) 1130 (610) N/A 951 (511) Steam temperature before desuperheating - F ( C) 889 (476) Steam production - lb/hr (kg/hr) 150,000 (68,039) Saturated steam temperature (t1) - F ( C) 500 (260) Boiler outlet gas temperature (t2) - F ( C) 545 (285) Pinch temperature (t2 - t1) - F ( C) 45 (25) Economizer outlet water temperature (t3) - F ( C) 467 (242) 730 (388) 67,000 (30,391) 495 (257) 517 (269) 22 (12) 495 (257) 812 (433) 118,500 (53,751) 497 (258) 537 (281) 40 (23) 491 (255) Approach temperature (tl - t3) - F ( C) 33 (18) 0 6 (3) Stack temperature - F ( C) 382 (194) 430 (221) 406 (208) NOTES: 1) Steam pressure at plant header is 625 prig (43.1 barg). 2) Feedwater temperature is 240 F (116 C) to economizer. 3) Steam temperature after desuperheating is 750 F (399 C) in both fired cases. this manner, the economizer was essentially split into two smaller economizers. By means of the arrangement of valves shown in Figure 3 it is possible to bypass water around the last four rows of the economizer. This gives an effective 40% reduction of economizer heating surface at low loads and prevents any tendency for steam formation to occur in the economizer. BACK PRESSURE Several design features are utilized to limit the backpressure to the 10 inches (254 mm) of water specified. In designing the heat exchangers, many different configurations were tried. These were judged from the standpoint of pressure drop, physical size, effectiveness of the heating surface and cost. The physical size of any heat exchanger can be characterized by two dimensions, frontal area and depth. Frontal area in this instance is the flow area for the exhaust gases. By varying this area, there is a corresponding increase or decrease in the gas side velocity and in the heat transfer coefficient. While frontal or flow area is important for both sides of the heat exchanger, in this application the gas side is more critical. This is because the relatively low heat transfer coefficients on the gas side are the major determinant in the overall heat transfer coefficient, the pressure drop on the gas side is much more critical than on the water/steam side and the designer has greater flexibility to change flow arrangement, circuitry, etc., on the water/steam side than the gas side. By increasing the frontal area, the velocity and pressure drop are decreased. However, the depth of the heat exchanger increases. This is due to the fact that the heat transfer coefficient decreases and therefore the heat transfer surface required for a given duty must increase. As the pressure drop is reduced, the weight, size and cost increases. The challenge then is to find a suitable compromise between size and pressure drop. At the same time, other physical limitations must be weighed. These include availability of materials, suitability of the design for the plant layout and ability to fabricate the equipment. In this instance the boiler size was first chosen. It was designed to have the largest physical size which could be shipped by railroad with a depressed center car. The physical characteristics of the heating surface within this envelope were then chosen to provide 5
a suitable compromise between minimizing pressure drop and providing acceptable performance. The superheater and economizer cross sections were then matched to the boiler. By the measures outlined above, pressure drop in the superheater, boiler and economizer were minimized. However, without additional measures to prevent excessive pressure loss in the ductwork, the maximum of 10 inches (254 mm) water column would still have been exceeded. Duct C (where the exhaust gases make the 90 degree turn from horizontal to vertical flow), was the place with highest possible pressure loss in the duct system. This duct was therefore provided with turning vanes to minimize the loss. These can be seen in Figure 1. The vanes were built entirely from stainless steel and equipped with provisions to accommodate the large scale thermal expansions that occur on startup. DE SUPERHEATER As mentioned previously, a desuperheater is required to reduce the steam temperature when the burner is in operation. Typically desuperheating is accomplished by spraying water into the stream of superheated steam. The energy required to vaporize this steam produces a corresponding decrease in the steam temperature. High purity desuperheating water is required. Because of the fact that the desuperheating water is sprayed directly into the steam, all suspended and dissolved solids in the desuperheating water are entrained in the superheated steam. In this plant, water of high enough quality was in short supply, so that a spray desuperheater was not attractive from this standpoint. A shell and tube heat exchanger would address this problem. With a heat exchanger there would be no direct mixing of the superheated steam and the cooling water and an ordinary source of cooling water could be used. This solution also presents its own problem which was also a concern of the user. As the HRSG cycles from the unfired to the fired case the desuperheater cycles from an inactive to an active mode. A potential problem is that during the inactive period steam contained in the heat exchanger would condense. Upon activation of the desuperheating exchanger a liquid slug of this condensed steam could be introduced into the steam turbine with disastrous results. A unique solution which addressed both of these concerns was adopted. A U-tube exchanger was placed in the lower drum of the boiler. When desuperheating is required, a portion of the superheated steam is passed through the desuperheater, overcooled and mixed back together with the non-desuperheated steam. Superheated steam is contained within the U-tubes. In the process of desuperheating, boiling occurs on the outside of the tubes. This method of desuperheating means that ordinary boiler water, which is required in any event, is used as the cooling medium. This has the additional advantage that the energy extracted from the superheated steam is returned to the system and is not wasted. Placing the exchanger in the lower drum of the boiler also prevents any condensation from occuring in the 20-0(6.1 M)--- ---{ A - LOWER BOILER DRUM INLET HEADER- (-30 (760 MM) I.D. I, -"U-TUBES" SLOPED FOR VENTING AND DRAINING MANWAY A J STEAML OUTLET HEADER INLET STEAM ' - FLANGE JOINT TO ALLOW OUTLET REMOVAL OF DESUPERHEATER SECTION A- A DESUPERHEATER ASSEMBLY FIGURE 4 exchanger. An examination of the fired base case operating data reveals why this is true. Steam pressure at the plant header is 625 psig (43.1 Barg). Allowing for a 5 psi (0.3 bar) drop from the superheater outlet to the plant header puts the superheater outlet pressure at 630 psig (43.4 barg). Saturation temperature at 630 psig (43.4 barg) is 494 degree F (257 degree C). Since the superheater outlet and the desuperheater are essentially at the same pressure, any steam in the desuperheater would have to be cooled below 494 degree F (257 degree C) for condensation to occur. Neglecting the small effect of static head, the pressure of the saturated water in the lower drum of the boiler is 666 psig (45.9 barg). This is because the superheater was designed for a 36 psi (2.5 bar) steam side pressure drop to assure good flow distribution into the many parallel steam circuits. Saturation temperature at 666 psig (45.9 barg) is 500 degree F (260 degree C). Therefore the steam in the desuperheater can never be cooled below 500 degree F (260 degree C) and will always remain in a slightly superheated state even during those inactive periods when the desuperheater is not operating. Figure 4 shows the mechanical details of the desuperheater. As can be seen, it consists of a series of U-tubes welded into inlet and outlet headers. Tubes are 2-1/2 inch (63.5 mm) outside diameter,.200 inch (5.1 mm) minimum wall low alloy carbon steel. Headers are special 9 inch (229 mm) outside diameter extruded pipe of 1-1/4% Cr, 1/2% Mo material. The headers are welded into a cylinder with elliptical head, manway and flange to mate to the lower drum. The tubes are supported over their length by rails which slide on tracks in the drum allowing the unit to be slid in or out. The U-tubes are sloped so that they can be vented and drained through the exterior piping. The inlet header is provided with a sleeve of Inconel 600. An air gap between the sleeve 00 and header ID provides insulation to help keep the header cool, reducing thermal stresses where it penetrates the shell. SUPPLEMENTARY BURNER The supplementary burner consisted of a standard grid type burner. This burner uses the high oxygen content 6
Li (approximately 15% by volume) of typical gas turbine exhaust as combustion air. Since ambient air need not be added and the turbine exhaust is already preheated, the supplementary firing is a highly efficient combustion process. The burner, as installed, was equipped for gas firing. Provision and space are included for the future addition of oil guns. Maximum duty of the burner is 88.6 million BTU/hr (26.0 MW). The design of the burner for gas firing is relatively simple. The burner consists of a section of ductwork. Mounted in this duct are four horizontal elements. Each element consists of a tubular gas riser with a series of evenly spaced orifices facing downstream. Attached to the riser at each orifice location is a cup shaped flame retainer. Each flame retainer and orifice acts as a small individual burner. In this manner a substantially uniform flame pattern and temperature distribution are obtained downstream of the burner. The flame retainer acts to stabilize the flame in a relatively protected area before the flame emerges into the main flow of the exhaust gases. Each of the four burner elements has an individual ignitor and ultraviolet flame scanner. The individual gas risers penetrate the duct wall and terminate in a header. Other items necessary to complete the burner package include the piping train, controls and flame safety system. 1/4"(6 MM) THK. GAS TIGHT STRUCTURAL OUTER CASING STIFFENER 12 GA.(2.8 MM) LINER " I/q'(6 MM) THK. LAP STRIP 1/2 (13 MM) 0 BOLT 2 STAGGERED LAYERS (304 SS) OF INSULATION TYPICAL CASING CONSTRUCTION DETAIL-SECTIONAL VIEW FIGURE 5 I/4(6MM)THK. CAI LAP STRIPS IUI l Critical to the successful operation of the supplementary burner is the ductwork adjacent to it. The burner requires a substantially even velocity profile across it. Because the velocity distribution in the gas turbine exhaust stream is typically uneven, special provisions are made to straighten the flow. A ^ perforated grid is provided immediately upstream of the burner. By sizing the grid for a pressure loss of about one velocity head, an even velocity profile can be obtained. Also, the ductwork is normally arranged so the burner has a considerable length of duct both upstream and downstream with a constant cross section. As can be seen in Figure 1, this was not possible in this case. This installation is normal in that a divergence in the duct cross section is required between the gas turbine exhaust plenum and the inlet to the HRSG. What is not normal is the fact that because of the limited space the burner had to be placed in the middle of this diverging section. Figure 2 shows the means taken to provide for proper burner operation. Upstream of the burner are five turning vanes to help the exhaust gas follow the duct divergence. These are located in the duct identified on the drawing as Duct A. The leading edge of the vanes was pivoted to provide a means for final field adjustment of the vanes. Located downstream of the vanes is the perforated grid of approximately 60% open area. The burner is then located downstream of this grid and the flame develops in Duct B. The length of Duct B was chosen to avoid flame impingement on the turning vanes located in Duct C. 12 GA. (2.8 MM) LINER IRS oi/2 (I3 MM) 0 BOLT (304 SS) HL J I LINER PANELS FREE I I I TO EXPAND IN ALL I, SLOTTED HOLES FOR DIRECTIONS Ii THERMAL EXPANSION TYPICAL CASING CONSTRUCTION DETAIL- ELEVATION VIEW FIGURE 6 casing must be engineering to accommodate the thermal expansions associated with this type of service. Refractories typically used in boiler construction are not adequate and will fail due to spalling occuring during the heatup. Additionally, the casing must be pressure tight and was specified to withstand an internal pressure of 20 inches (508 mm) water column. RAPID STARTUP In the design and fabrication of this or any other HRSG for use with a gas turbine, the construction of the boiler casing is of critical importance. This is due to the extremely rapid startup sequence and potentially cyclical nature of gas turbine operation. The As can be seen in the sectional view of Figure 5, the casing consists of a gas tight envelope of carbon steel plate. Insulation is applied to the inside of the casing. A liner of either stainless or carbon steel sheet is placed over the insulation to prevent it from being ripped away by the turbulent flow of exhaust gases. The concept of internally insulating - 7 -
I- the casing is extremely important. Since the casing is the pressure containing envelope, the required stiffeners and structural steel supports are welded to the outside of it. Internal insulation keeps this pressure containing envelope and supporting structure relatively cool. This reduces the thermal expansion to levels that can be easily dealt with. The thermal expansion must then be accommodated within the liner. By referring to Figure 5 and Figure 6 it can be seen how this is done. The inner liner is broken into small panels approximately 3 x 5' (900 mm x 1500 mm). Two methods are used to fasten this inner liner. In the interior regions of the small panels the liner is bolted to the casing by means of shoulder bolts. Oversize holes are backed up by large washers on each side. Nuts are threaded and tacked onto the bolts. In this way the liner is free to expand and contract. Large spaces are left between the edges of adjacent liner panels. These spaces are covered by lap strips of the same material as the liner. The lining panels are then free to expand and contract by sliding under the lap strips. With this means of fastening and the use of proper shoulder bolt spacing, the liner is prevented from being torn away by the turbine exhaust gas flow. FLOW VIBRATION Large heat recovery steam generators have occasionally been troubled by vibration problems. In the design of this HRSG a careful analysis of potential vibration problems was made. Based on this analysis means were provided to prevent these potential problems. There are two main sorts of vibration associated with these types of heat exchangers. They are vortex induced acoustic resonance and whirling instability of the tubes. Associated with tubes in crossflow there is a periodic vortex shedding which occurs. The vortex shedding frequency is a function of tube diameter, tube spacing and flow velocity over the tubes. The vortex shedding frequency can be predicted based on these parameters. The heat exchanger box can be considered as a cavity. As with any cavity, the acoustic natural frequency of this box can be predicted. When the vortex shedding frequency and the cavity frequency coincide, a vortex induced acoustic resonance may result. This produces a very loud, low frequency noise. While it is not generally damaging, it is undesirable. This can be avoided by tuning the cavity frequency. By the addition of baffles the characteristic dimension of the cavity can be reduced. This increases its acoustic natural frequency. By this means the acoustic natural frequency of the cavity can be raised above the vortex shedding frequency. In the case of this particular unit it was determined that baffles were required to tune the cavity of each heat exchanger. One baffle each was located in the superheater and economizer. Because of its larger characteristic dimension, a total of 5 baffles were installed in the boiler. Whirling instability of a bank of heat exchanger tubes in crossflow occurs when the flow velocity over the tube exceeds a certain value identified as the critical velocity. Whirling instability causes the tubes to vibrate in elliptical orbits. There is a synchronization between the orbits of the various individual tubes in the bundle. Whirling instability can be extremely damaging to the tubes. They may fail at their attachments to headers due to fatigue. There can also be a failure due to fretting wear of the tubes as they impact against their supports, the exchanger shell or each other. While the methods for predicting whirling instability are quite theoretical, design procedures to avoid this problem are known. Basically, the critical velocity increases with the natural frequency of the tubes. The designer can increase the natural frequency of the tubes by providing intermediate tube supports in order to reduce the free or unsupported length. In this manner the critical velocity for the onset of whirling instability can be raised above any expected operating velocities. An analysis of this HRSG system revealed that three intermediate tube supports were required in the superheater as well as in the economizer. None were required in the boiler section other than the integral baffle which is necessary for dividing the two passes. INSTALLATION AND OPERATIONAL EXPERIENCE Installation proceeded smoothly. The modular design helped to minimize erection time. The superheater, boiler and economizer as well as the ductwork arrived at the jobsite completely assembled. Major site work first consisted of foundation preparation. The structural steel assembly and module setting proceeded from the turbine exhaust flange outward through Duct C. The HRSG was then assembled from Duct C upwards. Once the major assemblies were set in place remaining site work consisted primarily of installation of the piping and walkways and ladders and insulation of piping. Startup proceeded with no major difficulties. After several months of operation, some fabrication errors were found in the installation of the turning vanes in Duct A. The owner chose to remove the turning vanes during a regularly scheduled turbine outage. No further operational or performance problems have been experienced. 8