Only terse answers are provided here. Interested readers may see my book "Steam Plant Calculations Manual" (page number where more information is given is shown for many questions) and also the Waste Heat Boiler Deskbook for detailed discussions,examples.

1.Relation between quality and purity is simple. Steam quality=100-(0.5/1000)x100=99.95 %

where 0.5 is the steam purity and 1000 ppm is the drum solids (p20).

2.Using the MM Btu method,quick estimates can be made for air required. For natural gas, air requires about 730 lb/MM Btu (Higher heating Value basis) and for fuel oil it is about 745. Hence at 10 % excess air,1 MM Btu gas fired requires 730x1.1=803 lb of dry air (p69) .

3.The relation between efficiency on HHV and LHV basis is as follows:

E

Typical natural gas has a HHV=23,000 Btu/lb and LHV=20,800 Btu/lb,with a ratio=1.105. Hence if HHV efficiency=83 %,the LHV efficiency=83x1.105=91.7 %. Boiler engineers should understand the difference while evaluating bids. Europeans typically use the LHV basis,while in the USA,HHV basis is the norm.

4.In an economizer,the water side heat transfer coefficent is 100 times higher than the gas side coefficient and hence the tube wall temperature will be close to the water temperature. Hence it does not matter if gas temperature is 600 F or 300 F,the tube wall temperature will be close to the water temperature. See my article on Corrosion in Economizers (p96).

5.The energy absorbed by steam in Btu/h is given by the equation Q=33475xbhp=33475x1000=33.475 MM Btu/h in a 1000 hp.boiler. From steam tables at 150 psig,enthalpy of vapor=1195.7 and that of feed water at 220 F=188.7 Btu/lb. Hence the steam generated= 33.475x10

6.See my article on Converting NOx,CO from mass to volumetric basis for the procedure to convert from mass to volume basis or vice versa. For natural gas,0.1 lb/MM Btu NOx=83 ppmvd. hence 56 ppmvd=0.0675 lb/MM Btu fired on HHV basis (p98) .

7.The heat loss from the casing in Btu/h can be shown to be nearly same if the ambient temperature,wind velocity are the same. However due to the lower emissivity (0.15 for aluminium vs 0.8-0.9 for steel),the aluminium casing will run hotter.The program on insulation performance may be used to verify this (p335).

8.This is an iterative process. Given the flue gas flow,inlet temperature,stack dimensions,one can estimate the casing heat loss,compute the inside heat transfer coefficient and the drop in gas temperature for a given stack/duct height.See Books,Software on Boilers,HRSGS,Steam Plant Calculations for information on a program to perform this calculation

(p 333).

9.Fans for boilers should be sized at the lowest density conditions or highest ambient/elevation conditions. This is due to the fact that air for generating a given amount of steam in a boiler in lb/h is nearly constant for a given fuel input. However the volume will be higher at lower densities and fans have to be sized to handle this (p363).

10.It can be shown that for a fan there is a relation between flow,head developed and motor current.

P=1.17x10-4xqH/h

P=power consumption of fan,kw

q=flow of air,acfm

H=head developed,in wc

E,I=motor voltage,current

cosF =power factor

h

Thus knowing current,voltage,head developed,one can estimate the flow. Some iteration may be required to adjust for fan efficiency,which varies with load. Simialr expression is available for pumps (p381).

11.The adiabatic combustion temperature is obtained by equating the net heat released by combustion of the fuel with the product of flue gas quantity generated and the enthalpy increase.

say fuel oil having a HHV=19,700 and LHV=18,500 Btu/lb is fired with 10 % excess air.The flue gas generated using MM Btu method is: 1+745x1.1x19700/10

12.For a deatiled expalnation,see the book(p195).For bare tubes,inline arrangement gives the lowest gas pressure drop for comparable duty and surface area. The heat transfer coefficients between inline and staggered arrangements is not very signifcant ,on the order of 5%;however the gas pressure drop is much higher,about 30-50 % more for staggered over inline. Hence use inline arrangements for bare tubes.On the other hand,for finned tubes,both arrangements are feasible and the differences are not very significant for the same duty and gas pressure drop.Cost,past practice and experience more often determine the arrangement.

13.A scale of given thickness will cause more problems in a finned tube boiler than in a bare tube boiler. The heat flux inside finned tubes is very high as seen in the article Heat Transfer in Finned Tubes . If the overall heat transfer coefficient =7 Btu/ft

For a bare tube boiler,the overall heat transfer coefficient is say 13 Btu/ft

14.True. A tube subject to external pressure reuqires higher thickness to withstand it.See ASME code for formule.Typically the pressure can be twice if applied inside comapred to that applied externally.(p41)

15.True. One has to develop the head vs flow curve for the system and see where it intersects the system resistance to determine the operating point and it may not,depending on the nature of the curve deliver the desired flow. Same applies to pumps(p382)

16.It can be shown from fundementals(p103) that the relation between fuel input (natural gas,distillate oils) and oxygen is:

17.More surface area does not mean more duty as discussed in the article on finned tubes. Even with bare tubes,depending on tube spacing ,gas velocity,tube size etc,the heat transfer coefficient can vary.

18.The article on finned tubes(see above Q 13) also shows how with better choice of fins,more energy can be transferred with lesser surface!

19.Higher fin density or larger ratio of external to tube internal surface area is preferred when the tube side heat transfer coefficient is large,as in the case of evaporators. When tube side coefficient is small,as in the case of superheaters,the large fin surface adds little to the duty but contributes negatively to heat flux,gas pressure drop and tube wall temperatures and hence should be avoided. See the example in the book (p254).

20.True. Lower the tube size,higher the heat transfer coefficient and shorter the tube length,though the labor cost may be higher as more tubes have to be used (p214).

21.See the article Converting NOx,CO emissions from mass to volumetric units for examples and procedure.

22.A higher Circulation Ratio does not mean larger duty for the evaporator. The duty depends on heat transfer coefficient,surface area and log-mean temperature difference.

23.Fouling caused by the scale=0.1/10=.01 ft

24.One cannot arbitrarily select the exit gas temperature in a gas turbine HRSG,due to limitations of pinch and approach points,which affect the exit gas temperature. See HRSG simulation for examples,discussions on this subject.

25.Gas turbine HRSGs have some special characteristics. The exhaust gas flow increases at lower ambient temperature but the exhaust gas temperature decreases. This combination generates less steam in the evaporator,which means less flow through the economizer. However the gas side heat transfer coefficient,which determines the overall heat tarnsfer coefficient,is the same or even higher at the economizer compared to the high ambient temperature or load case.This causes more energy to be transferred to the economizer. With lesser flow and more duty,the enthalpy absorbed is higher and water exit temperature approaches the saturation temperature and hence can lead to steaming. The same situation arises when load of the gas turbine decreases. The exhaust gas temperature is lower,while the mass flow is unchanged. Again less steam is generated,more enthaply absorbed in economizer and so steaming is a possibility.

Hence methods to avoid steaming include bypassing of the gas across the HRSG or the economizer (lower gas flow helps to reduce the energy transferred to the economizer),increasing the flow of water by recirculation or even by using higher blow down. See the Waste Heat Boiler Deskbook for more deatiled discussions.

26.The gas side heat transfer coefficient is lower as the fin density increases for the same gas velocity,temperature. See the article on Finned Tubes above.

27.Yes. A packaged steam generator can have different combinations of furnace area,convection surface and economizer surface for the same duty,gas pressure drop. Hence one should not simply go by surface areas alone.

28.A 100,000 lb/h steam generator handles a flue gas of about 110 to 130,000 lb/h depending upon the excess air and duty,assuming no flue gas recirculation. This results in a power consumption of about 4 kw/in wc of pressure drop. Cost conscious plant engineers must evaluate cost of fan operation along with fuel consumption. It is possible that a low gas pressure drop design is better in the long run even if the cost is slightly higher than a high pressure drop design.

29.A 10 % increase in surface area does not translate into 10 % more duty. Depending upon how the surface is added(by increase/decrease in gas velocity),the change could be from 2 to 5 %. As more energy is transferred,the log-mean-temperature-difference decreases. Hence more S does not mean more Q even if U is unchanged.Note the familiar equation Q=USDT.

30.Finned tubes make the design of water tube boilers compact as seen in the article above on Finned Tubes.

31.Multiple pressure steam generation is necessary in gas turbine HRSGS as the exit gas temperature cannot be decreased by a single pressure system particularly in unfired units. As seen in Generating Steam Efficienctly in Cogeneration Plants the exit gas temperature in unfired mode is higher.Exit gas temperature increases as the steam pressure and steam temperature increase(p286). With 15 F pinch and 20 F approach,the exit gas temperature in a 600 psig,750 F HRSG is 398 F,while the exit gas temperature in a 150 psig saturated unit is 313 F. This is a thermodynamic problem. Hence in high pressure units,multiple pressure modules are required to lower the exhaust gas temperature. One can run the HRSG simulation program to obtain these numbers. HRSGS program may also be used to determine if multiple pressures are necessary.

32.It is desirable to extract as much energy from the exhaust gas as possible.Hence feed water heating in the HRSG is a good idea. However one has to evaluate potential corrosion concerns if water vapor condensation or acid condensation is likely.Note that the water temperature entering detrmines if condensation is likely as shown in the article above on corrosion in economizers.

33.As shown in the article above on generating steam efficienctly in cogeneration plants,it is more efficient to add the fuel in the HRSG rather than in the steam generator.

34.The exit gas temperature increases with decrease in steam flow. That is due to the larger ratio of gas to steam in the HRSG.This is also responsible for steaming in economizers.

35.True. Hydrogen and water vapor increase the gas heat transfer coefficient by increasing the gas specific heat and thermal conductivity.A high gas pressure also increases the mass velocity,resulting in higher heat transfer coefficient as in hydrogen plant reformed gas boilers.

36.Yes. See the HRSGS simulation web site.

37.With air heaters,the air and gas side heat transfer coefficients are of the same order..Hence it makes no sense to use fins either inside or outside. See Q 19.Fins are recommended when the tube side coefficient is several times the gas side coefficient as in the case of economizers,evaporators and to some extent the superheater.

38.Steam generators at low loads have several problems. First it is difficult to predict the performance due to poor gas/steam side mal distrubution. The fan also may operate at an unstable point,unless variable speed drives are used. If radiant superheaters are used,they are susceptable to failures due to low tube side pressure drop and poor tube side flow distribution and associated higher radiant energy transfer.Good performance prediction can generally be made between 60 to 100 % load.

39.Typical exhaust gas temperature is 850-1000 F. HRSGs can be designed economically for a pinch and approach point of 10-20 F for these temperatures and the size/surface areas will be reasonable and feasible. In fired conditions,pinch and approach points cannot be arbitrarily selected as it can lead to steaming in the economizer in unfired mode or temperature cross situation can arise at other fired conditions. If the HRSG is sized in unfired mode,its performance at other loads,both fired or low loads can be evaluated using the simulation process to assure soundness of design.See my book Waste Heat Boiler Deskbook for elaborate discussions on pinch,approach point selection.

40.See the section on Fouling in Waste Heat Boilers

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