01 POWER ISLAND / 01 CCPP / V. Ganapathy-Industrial Boilers and HRSG-Design (2003)

TABLE 8.30 Performance of Alternative HRSG Designs

Gas flow ¼ 287,000 lb=h; Gas analysis (vol%) CO2 ¼ 3, H2O ¼ 7, N2 ¼ 75, O2 ¼ 15. Steam pressure ¼ 300 psig sat; gas turbine power ¼ 8000 kW.

Due to higher gas pressure drop of 1.3 in.WC:

1:3 8000 0:07 1008000 4 ¼ $14;560

Thus the net benefit of using design A over B is $(33,240 þ 46,480 7 14,560) ¼ $65,160 per year.

If the additional cost of design A over B due to its size is, say, $50,000, the payback of using design A is less than 1 year. However, if the HRSG operates for less than, say, 3000 h=year, the payback will be longer and has to be reviewed.

8.41

Q:

What is steaming, and why is it likely in gas turbine HRSGs and not in conventional fossil fuel fired boilers?

A:

When the economizer in a boiler or HRSG starts generating steam, particularly with downward flow of water, problems can arise in the form of water hammer, vibration, etc. With upward water flow design, a certain amount of steaming, 3– 5%, can be tolerated because the bubbles have a natural tendency to go upward along with the water. However, steaming should generally be avoided.

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TABLE 8.31 Typical Exhaust Gas Flow, Temperature Characteristics of a Gas Turbine

Fuel: natural gas; elevation: sea level; relative humidity 60%; inlet loss 4 in.H2O; exhaust loss 15 in.H2O; speed: 3600 rpm; output terminal: generator.

To understand why the economizer is likely to steam, we should first look at the characteristics of a gas turbine as a function of ambient temperature and load (see Tables 8.31 and 1.4).

In single-shaft machines, which are widely used, as the ambient temperature or load decreases, the exhaust gas temperature decreases. The variation in mass flow is marginal compared to fossil fuel fired boilers, while the steam or water flow drops off significantly. (The effect of mass flow increase in most cases does not offset the effect of lower exhaust gas temperature.) The energytransferring ability of the economizer, which is governed by the gas-side heat transfer coefficient, does not change much with gas turbine load or ambient temperature; hence nearly the same duty is transferred with a smaller water flow through the economizer, which results in a water exit temperature approaching saturation temperature as seen in Q8.35. Hence we should design the economizer such that it does not steam in the lowest unfired ambient case, which will ensure that steaming does not occur at other ambient conditions. A few other steps may also be taken, such as designing the economizer [8] with a horizontal gas flow with horizontal tubes (Fig. 8.17). This ensures that the last few rows of the economizer, which are likely to steam, have a vertical flow of steam–water mixture.

In conventional fossil fuel fired boilers the gas flow decreases in proportion to the water flow, and the energy-transferring ability of the economizer is also lower at lower loads. Therefore steaming is not a concern in these boilers; usually the approach point increases at lower loads in fired boilers, whereas it is a concern in HRSGs.

The other measures that may be considered to minimize steaming in an economizer are

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FIGURE 8.17 Horizontal gas flow economizer.

Increase the water flow through the economizer during these conditions by increasing the blowdown flow. This solution works only if small amounts of steam are formed and the period of operation in this mode is small. Blowdown results in a waste of energy.

Increasing the inlet gas temperature either by supplementary firing or by increasing the turbine load helps to generate more steam and thus more water flow through the economizer, which will prevent steaming. As we saw in Chapter 1, the economizer steams at low loads of the turbine.

Exhaust gases can be bypassed around the HRSG during such steaming conditions. This minimizes the amount of energy transferred at the economizer as well as the evaporator. Gas can also be bypassed around the economizer, mitigating the steaming concerns.

Water can also be bypassed around the economizer during steaming conditions, but this is not a good solution. When the gas turbine load picks up, it will be difficult to put the water back into the economizer while the tubes are hot. The cold water inside hot tubes can flash and cause vibration and thermal stresses and can even damage the economizer tub.

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8.42

Q:

Why are water tube boilers generally preferred to fire tube boilers for gas turbine exhaust applications?

A:

Fire tube boilers require a lot of surface area to reduce the temperature of gas leaving the evaporator to within 15–25 F of saturation temperature (pinch point). They have lower heat transfer coefficients than those of bare tube water tube boilers (see Q8.10), which do not compare well with finned tube boilers. Water tube boilers can use extended surfaces to reduce the pinch point to 15–25 F in the unfired mode and hence be compact. The tubes will be very long if fire tube boilers are used; hence the gas pressure drop will be higher. (A fire tube boiler can be made into a two-pass boiler to reduce the length; however, this will increase the shell diameter and the labor cost, because twice the number of tubes will have to be welded to the tube sheets.) The fire tube boiler will have to be even larger if the same gas pressure drop is to be maintained. Table 8.32 compares the performance of water tube and fire tube boilers for the same duty and pressure drop.

It can be seen from the table footnotes that the water tube boiler is very compact. If the gas flow is very small, say less than 50,000 lb=h, then a fire tube boiler may be considered.

TABLE 8.32 Water Tube vs. Fire Tube Boiler for Gas Turbine

Exhaust

aWater tube boiler: 2 0.105 in. tubes, 20 wide, 18 deep, 6 ft long, with 5 serrated fins=in., 0.75 in. high, 0.05 in. thick.

bFire tube boiler: 1400 1.5 0.105 in. tubes, 21 ft long.

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8.43

Q:

Does the addition of 10% surface area to a boiler increase its duty by 10%?

A:

No. The additional surface area increases the duty only slightly. The increased temperature drop across the boiler and the temperature rise of water or steam (if single-phase) due to the higher duty results in a lower log-mean temperature difference. This results in lower transferred duty, even assuming that the overall heat transfer coefficient U remains unchanged. If the larger surface area results in lower gas velocities, the increase in duty will be marginal as U is further reduced.

As an example, consider the performance of a fire tube boiler with 10% and 20% increase in surface area as shown in Table 8.33. As can be seen, a 10% increase in surface area increases the duty by only 3%, and a 20% increase in surface area increases the duty by only 6%. Similar trends may be shown for water tube boilers, superheaters, economizers, etc.

8.44a

Q:

How do we estimate the time required to heat a boiler?

A:

A boiler can take a long time to heat up, depending on the initial temperature of the system, mass of steel, and amount of water stored. The following procedure gives a quick estimate of the time required to warm up a boiler. The methodology is applicable to either fire tube or water tube boilers.

TABLE 8.33 Boiler Performance with Increased Surface Areaa

aGas flow ¼ 70,000 lb=h; inlet gas temperature ¼ 1600 F. Gas analysis (vol%): CO2 ¼ 7, H2O ¼ 12, N2 ¼ 75, O2 ¼ 6; steam pressure ¼ 125 psig saturated. Tubes: 2 0.120 carbon steel.

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Gas at a temperature of Tg1 enters the unit, which is initially at a temperature of t1 (both the water and the boiler tubes). The following energy balance equation can then be written neglecting heat losses:

where

Mc ¼ water equivalent of the boiler

¼mass of steel specific heat of steel þ mass of water specific heat of water (Weight of the boiler tubes, drum, casing, etc., is

included in the steel weight.)

dt=dz ¼ rate of change of temperature, F=h Wg ¼ gas flow, lb=h

Cpg ¼ gas specific heat, Btu=lb F

Tg1; Tg2 ¼ entering and exit boiler gas temperature, F U ¼ overall heat transfer coefficient, Btu=ft2 h F A ¼ surface area, ft2

DT ¼ log-mean temperature difference, F

¼ ðTg1 tÞ ðTg2 tÞ ln½ðTg1 tÞ=ðTg2 tÞ&

t ¼ temperature of the water=steam in boiler, F

From Eq. (81) we have

"#

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To estimate the time to heat up the boiler from an initial temperature t1 to t2, we have to integrate dt between the limits t1 and t2.

The above equation can be used to estimate the time required to heat the boiler from a temperature of t1 to t2, using flue gases entering at Tg1. However, in order to generate steam, we must first bring the boiler to the boiling point at atmospheric pressure and slowly raise the steam pressure through manipulation of vent valves, drains, etc; the first term of Eq. (81) would involve the term for steam generation and flow in addition to metal heating.

Example

A water tube waste heat boiler of weight 50,000 lb and containing 30,000 lb of water is initially at a temperature of 100 F. 130,000 lb of flue gases at 1400 F enter the unit. Assume the following:

Gas specific heat ¼ 0.3 Btu=lb F

Steel specific heat ¼ 0.12 Btu=lb F

Surface area of boiler ¼ 21,000 ft2

Overall heat transfer coefficient ¼ 8 Btu=ft2 h F

Estimate the time required to bring the boiler to 212 F.

or z ¼ 0.084 h ¼ 5.1 min.

One could develop a computer program to solve Eq. (81) to include steam generation and pressure-raising terms. In real-life boiler operation, the procedure is corrected by factors based on operating data of similar units.

It can also be noted that, in general, fire tube boilers with the same capacity as water tube boilers would have a larger water equivalent and hence the start-up time for fire tube boilers would be longer.

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8.44b

Q:

Assuming that the superheater in Q8.19c is dry, how long does it take to heat the metal from 80 F to 900 F? Assume that the gas-side heat transfer coefficient is 12 Btu=ft2 h F. Gas flow and temperature are the same as before. The weight of the superheater is 5700 lb. 150,000 lb=h of exhaust gases enter the superheater at 1030 F.

"#

z ¼ 0:0737 h ¼ 4:5 min

This is an estimate only but gives an idea of how fast the metal gets heated up. This is important in gas turbine plants without a gas bypass system. A large quantity of exhaust gases can increase the metal temperatures quickly. Hence if frequent start-ups and shutdowns are planned, a stress analysis is required to ensure that critical components are not subjected to undue stresses due to quick changes in tube wall or header temperatures.

By the same token, the superheater tubes cool fast when the exhaust gas is shut off compared to, say, evaporator tubes, which are still hot due to the inventory of hot saturated liquid. This can lead to condensation of steam when the HRSG is restarted, leading to blockage of flow inside the superheater tubes unless adequate drains are provided.

8.44c

Q:

A large mass of metal and water inventory in a boiler results in a longer start-up period, but the residual energy in the metal also helps to respond to load changes faster when the heat input to the boiler is shut off. Drum level fluctuations also are

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smoothed out by a large water inventory. In order to understand the dynamics, let us look at an evaporator in a waste heat boiler with the following data:

Gas flow ¼ 350,000 lb=h Gas inlet temp ¼ 1000 F Gas exit temp ¼ 510 F

Steam pressure ¼ 600 psig sat Feedwater temp ¼ 222 F

Tubes: 2 0.105, 30 tubes=row, 20 deep, 12 ft long with 4.5 0.75 0.05 in. serrated fins

Steam drum ¼ 54 in., mud drum ¼ 36 in; both are 13 ft long. Boiler generates 45,000 lb=h of steam.

Weight of steel including drums ¼ 75,000 lb Weight of water in evaporator ¼ 18,000 lb Volume of steam space ¼ 115 ft3

Feedwater temperature ¼ 220 F

Energy transferred by gas to evaporator ¼ 45.9 MM Btu=h

What happens to the steam pressure and steam generation when the heat input and the feedwater supply are turned off?

dh=dp ¼ change of enthalpy to change in pressure, Btu=lb psi

dT=dp ¼ change of saturation temperature to change in pressure, F=psi Q ¼ energy transferred to evaporator, Btu=h

dp=dz ¼ rate of pressure change, psi=h

Now assuming that the volume of space between the drum level and the valve ¼ V ft3, we can write the following expression for change in pressure using the perfect gas law:

pV ¼ C ¼ pV =m

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where

C ¼ a constant

m ¼ mass of steam, lb, in volume V

where

p ¼ pressure, psia

Ws; Wl ¼ steam generated and steam withdrawn, lb=h

For steam at 600–630 psia, we have from the steam tables that the saturation

The pressure decay will be about 1.59 psi=s if this situation continues without correcting feedback such as matching heat input and feedwater flow.

These calculations, though simplistic, give an idea of what happens when, for example, the turbine exhaust gas is switched off. In fresh air fired HRSGs, there is a small time delay, on the order of a minute, before the fresh air fired

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