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PROJECT REPORT ON

MECHANICAL DESIGN OF

HEAT RECOVERY STEAM GENERATOR

UNDERTAKEN AT

THERMAX LIMITED, PUNE, MAHARASHTRA

For the partial fulfilment of requirement of B.Tech degree in

Mechanical Engineering

Submitted By:

Anurag Baruah

B.Tech (Mech.)

2010BTME012

Shridhar University, Pilani, Rajasthan

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ABSTRACT

Designing and drawing are the most important aspects of mechanical engineering. Design

incorporates a detailed study and development of various components required for assembly

into final product. Successful assembly of the final product after manufacture depends to a

great extent upon the design of the components. The gist of design lies in the accuracy with

which the properties and behaviour of the product can be ascertained.

The objective of the project is to design a “HEAT RECOVERY STEAM GENERATOR

(HRSG)” so that the overall efficiency of the plant gets increased. The system comprises of a

high pressure steam drum, natural circulation, and unfired water tube boiler designed in

accordance with “INDIAN BOILER REGULATIONS (IBR)” and “AMERICAN SOCIETY

OF MECHANICAL ENGINEERS (ASME) SEC I & SEC VIII.” The exhaust from the gas

turbines enters the HRSG via a duct. The HRSG consists of a plain tube co-current drainable

super-heater section followed by modular sections of convection banks and counter current

drainable economiser banks. The exhaust gas from the gas turbines enters through the super

heater and is released into the atmosphere via a stack after in passes through the economiser

section. The gas loses its thermal energy during its flow and thus provides the heat to convert

water into steam.

This type of boiler mainly finds application in industries which use gas turbines. Using the

exhaust gas for production of steam increases the overall efficiency of the plant.

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CONTENTS

1. THERMAX LIMITED: AN OVERVIEW 6

2. INTRODUCTION TO BOILERS 7

3. INTRODUCTION TO HRSG 10

4. MECHANICAL DESIGN OF HRSG ACC. TO IBR 21

5. MECHANICAL DESIGN OF HRSG ACC. TO ASME 39

6. DESIGN CALCULATION 54

7. PROJECT INFERENCE 69

8. CONCLUSION 70

9. REFERENCE 71

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THERMAX LIMITED: AN OVERVIEW

Thermax Ltd. is an Indian energy and environment engineering company based in India and

Britain. It manufactures boilers, vapours absorption machines, offers water and waste

solutions and installs captive power projects.

HISTORY:

Thermax came into being by harnessing the power of steam, produced by boilers. The

company first started with producing small, once through, baby boilers for catering steam

required at that time by the hospitals. In 1966, it collaborated with a Belgian company,

Wanson, to commence business operations as Wanson India Ltd., manufacturing small boiler

at a unit in Dadar, Mumbai. The company was renamed Thermax Limited in 1980.

THERMAX LTD. (INDIA):

Company based in India, which has also made boilers, has four manufacturing centres, and

operates in seventy five countries. It became known as Thermax Limited in 1980. In 1987 it

started making vapour absorption machine, in collaboration with Sanyo of Japan. It formed a

joint venture in 1988 with North Carolina- based Babcock and Wilcox, who make boilers, to

make steam generation units for heat recovery steam generators (HRSGs). In 1992 it formed

its Combined Heat and Power Group.

STRUCTURE:

It has four main offices-

Thermax (Europe) Ltd is based in Fenny Stratfort, Milton Keynes in England (not far

from IKEA).

Thermax Inc – based in Nothville, Michigan, USA

Thermax do Brasil Energia e Equipamentos Ltda – Brazil

Thermax (Zhejiang) Cooling & Heating Engineering Co – China

Its main divisions are:

Cooling and Heating (C & H)division.

Boilers and Heaters (B & H)division.

Power division.

Enviro division.

Chemical and Water division.

Solar energy division.

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INTRODUCTION TO BOILERS

DEFINITION:

A boiler is a closed vessel in which water or other fluid is heated. The fluid does not

necessarily boil. The heated or vaporised fluid exits the boiler for use in various processes or

heating applications including central heating, boiler based power generation, cooking and

sanitation.

MATERIALS:

The pressure vessel of a boiler is usually made of steel (or alloy steel) or historically of

wrought iron. Stainless steels, especially of the austentic types, are not used in the wetted

parts of the boilers due to corrosion and stress corrosion cracking. However, ferritic stainless

steels are often used in superheater sections that will not be exposed to boiling water.

FUELS:

The source of heat for a boiler is combustion of any of several fuels, such as wood, coal, oil

or natural gas. Electric steam boilers use resistance- or immersion-type heating elements.

Nuclear fission is also used as a heat source for generating steam, either directly or in

specialised heat exchangers called “steam generators”. Heat Recovery Steam Generators use

the heat rejected from other processes such as gas turbines.

CLASSIFICATION OF BOILERS:

Boilers are usually classified on the following basis:

1. Based on utilization:

Industrial boilers: They are utilised to produce steam for electric power generation.

They normally have a large capacity, high steam parameters and high boiler

efficiency.

Marine boilers: They are used a source of motive power for ships. They normally

have a compact shape, lighter weight and mostly fuel oil fired..

2. Based on steam/ water circulation:

Natural circulation boiler: The circulation of the working fluid in the evaporating

tube is produced by the difference in density between the steam/ water mixture in the

risers and water in the downcomers.

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Forced multiple circulation boilers: The circulation of the working fluid in the

evaporating tube is forced by means of a circulating pump included in circulation

circuit.

Once through boilers: There is no drum. The working fluid passes through the

evaporating tubes only once under the action of feed water pump.

Combined circulation boilers: the system includes a pump, back pressure valve, and a

mixer in the circuit. At starting the back pressure valve is opened and the boiler

operates as a forced multiple circulation boiler.

3. Based on pressure:

Low to medium pressure boilers:

High pressure boilers:

Super-critical pressure boilers:

4. Based on heat source used:

Solid fuel fired boiler: These types of boilers are typically low cost. The components

of the fuel the characteristics of the ash are important factors for boiler design.

Fuel oil fired boiler: These types of boilers have higher flue gas velocity and smaller

furnace size.

Gas fired boiler: These types of boilers utilise natural gas with higher flue gas

velocity and smaller furnace volume.

Waste heat boiler: These types of boilers utilise waste heat from any industrial

process for heating purpose.

5. Based on tube layout:

Water tube boiler: It is a type of boiler in which water circulates in tubes fired

externally by fire. Fuel is burned inside the furnace, creating hot gas which heats

water in the steam generating tubes. In smaller boilers, additional generating tubes are

separate in the furnace, while larger utility boilers rely on the water-filled tubes that

make up the walls of the furnace to generate steam.

The heated water then rises into the steam drum. Here, saturated steam is drawn off

the top of the drum. In some services, the steam will re-enter the furnace through a

super-heater to become superheated. Superheated steam is defined as steam that is

heated above the boiling point at a given pressure. Superheated steam is a dry gas and

therefore used to drive turbines, since water droplets can severely damage turbine

blades.

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Cool water at the bottom of the steam drum returns to the feedwater drum via large-

bore 'downcomer tubes', where it pre-heats the feedwater supply. In large utility

boilers, the feedwater is supplied to the steam drum and the downcomers supply water

to the bottom of the waterwalls. To increase economy of the boiler, exhaust gases are

also used to pre-heat the air blown into the furnace and warm the feedwater supply.

Such watertube boilers in thermal power stations are also called “steam generating

units”.

The older fire-tube boiler design – in which the water surrounds the heat source and

the gases from combustion pass through tubes through the water space – is a much

weaker structure and is rarely used for pressures above 350 psi (2.4 MPa). A

significant advantage of the water tube boiler is that there is less chance of a

catastrophic failure: there is not a large volume of water in the boiler nor are there

large mechanical elements subject to failure.

Fire tube boiler: A fire-tube boiler is a type of boiler in which hot gases from a fire

pass through one or more tubes running through a sealed container of water. The

heat of the gases is transferred through the walls of the tubes by thermal conduction,

heating the water and ultimately creating steam.

The fire-tube boiler developed as the third of the four major historical types of boilers:

low-pressure tank or "haystack" boilers, flued boilers with one or two large flues, fire-

tube boilers with many small tubes, and high-pressure water-tube boilers. Their

advantage over flued boilers with a single large flue is that the many small tubes offer

far greater heating surface area for the same overall boiler volume. The general

construction is as a tank of water penetrated by tubes that carry the hot flue

gases from the fire. The tank is usually cylindrical for the most part—being the

strongest practical shape for a pressurised chamber—and this cylindrical tank may be

either horizontal or vertical.

This type of boiler was used on virtually all steam locomotives in the horizontal

"locomotive" form. This has a cylindrical barrel containing the fire tubes, but also has

an extension at one end to house the "firebox". This firebox has an open base to

provide a large grate area and often extends beyond the cylindrical barrel to form a

rectangular or tapered enclosure. The horizontal fire-tube boiler is also typical of

marine applications, using the Scotch boiler. Vertical boilers have also been built of

the multiple fire-tube type, although these are comparatively rare: most vertical

boilers were either flued, or with cross water-tubes.

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INTRODUCTION TO

HEAT RECOVERY STEAM GENERATOR

AN OVERVIEW ON HRSGs:

A heat recovery steam generator or HRSG is an energy recovery heat exchanger that recovers

heat from a hot gas stream. It produces steam that can be used in a process (co-generation) or

used to drive a steam turbine. The fundamental purpose of HRSG is to extract the useful

energy in waste heat, either from the exhaust of a gas turbine or reciprocating engine.

HRSGs consist of four major components: the economiser, evaporator, super-heater and

water pre-heater. The different components are put together to meet the operating

requirements of the unit. See the attached illustration of a Modular HRSG General

Arrangement.

Modular HRSGs can be categorized by a number of ways such as direction of exhaust gases

flow or number of pressure levels. Based on the flow of exhaust gases, HRSGs are

categorized into vertical and horizontal types. In horizontal type HRSGs, exhaust gas flows

horizontally over vertical tubes whereas in vertical type HRSGs, exhaust gas flow vertically

over horizontal tubes. Based on pressure levels, HRSGs can be categorized into single

pressure and multi pressure. Single pressure HRSGs have only one steam drum and steam is

generated at single pressure level whereas multi pressure HRSGs employ two (double

pressure) or three (triple pressure) steam drums. As such triple pressure HRSGs consist of

three sections: an LP (low pressure) section, a reheat/IP (intermediate pressure) section, and

an HP (high pressure) section. Each section has a steam drum and an evaporator section

where water is converted to steam. This steam then passes through super-heaters to raise the

temperature beyond the one at the saturation point.

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TYPES OF HRSGs:

1. Horizontal HRSG: It features natural circulation typically consisting of multi-pressure

steam systems- high pressure (HP), intermediate pressure (IP) and low pressure (LP).

Triple pressure units add a reheat system to further boost the overall cycle’s thermal

efficiency.

A horizontal HRSG typically is installed either in-line with the axis of the gas turbine or

perpendicular to that axis.

Benefits:

Modular design with much of the connecting pipe systems fabricated in the factory.

Modular units require a reduced number of field welds, hence their erection tends to

be faster and less prone to error.

Standardised modules in most designs, some with uniform headers and standardised

connecting piping.

High temperature alloy in super-heater and re-heater sections. The advanced alloy

typically 9% chrome-moly steel adds to strength, resistance to oxidation and

protection against creep and thermal fatigue.

Low allow steel for locations vulnerable to flow accelerated corrosion such as riser

piping to intermediate pressure (IP) & low pressure (LP) drums.

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2. Vertical HRSG: It is better suited than a horizontal design for cycling and load following

duties. The primary disadvantage of vertical HRSG is that they require forced circulation.

Newer vertical design, however, combine the load following durability of the vertical

design with the high efficiency natural circulation, requiring circulation pumps only

during transient periods.

Benefits:

Vent able serpentine shape tube arrangement that allow for free expansion.

A ‘warm casing’ design where the casing expand in unison with the tube bundle,

thus reducing thermal stress.

Smaller water inventory, hence, less thermal inertia.

Effective and accessible drainage system with the lowest pressure point being

approximately 20 ft above ground.

Simpler design with very few headers and inter-connecting pipes.

Overall lighter boilers.

SPECIALISED FEATURES OF HRSGs:

Flexibility of design: HRSGs are available in un-fired, supplementary fired and fresh air-

fired mode of operations. It also consists of multiple pressure level with reheat and integral

de-aerator.

Natural circulation design: Absence of external circulating pumps results in lower power

consumption and eliminates typical problems associated with high pressure and high

temperature re-circulating pumps. Also, natural circulation design ensures high reliability

and availability of units.

Single drum construction: Steam drum is located outside the gas path to reduce thermal

stress. This facilitates quicker starts and stops by the HRSG to follow gas turbine

operations.

Fully welded construction for single drum: All pressure parts are welded to headers,

thereby allowing quicker start- ups and shut- downs of the HRSG to follow gas turbine

operations.

Bi-drum construction: Bi-drum construction for gas turbines below 15MW capacity. In

this design the pressure part tubes are expanded into the steam drum and water drum.

Special drum internals: Specially designed drum internals installed in the steam drum

promote circulation and ensure supply of bubble free intaining high steam purity (99.99%

dryness). This results in stable performance during quick load pickups and reductions.

Carefully designed inlet duct: HRSG performance can be drastically affected by

misdistribution of turbine exhaust gas which is highly turbulent. They come with carefully

designed inlet duct, angle of transition duct and flow distribution grid thus achieving

HRSG performance and preventing overheating of super-heater tubes.

Tolerant to economiser steaming: The last rows of the economiser are arranged in an up-

flow configuration. These rows help avoid any imbalance in water distribution associated

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with steaming in the economiser lower loads. Water leaving the economiser enters the

baffled portion of the steam drum and passes through hydro-cyclone separators to further

improve water/steam flow distribution and circulation.

Radiation screening: The first two rows following the burner, in case of a supplementary

fired unit, are constructed of bare tubes arranged in staggered pitching formation. These

two rows screen flame radiation coming from the burner and help to avoid overheating of

fins in subsequent HRSG surface area.

Gas- tight internally insulated casing: The entire HRSG is enclosed in a gas- tight

casing with stiffeners provided to enhance rigidity. This design is based on the concept of

‘cold casing’ using ceramic mattresses. Specially designed studs hold ceramic wool

insulation material tightly to the outer casing. This results in minimised thermal expansion

of the casing and thermal loads on the gas turbine flange. For supplementary fired units,

pyroblocks are provided in the downstream duct after burner to avoid distortion of liners

due to high temperature.

Safe location of welds outside the gas path: All header- to- tube welds are located

outside the active gas path, thus avoiding direct contact of welds with high-velocity, high-

temperature flue gases. This enhances safety of all pressure part joints. For super-heaters,

header protection is provided to limit metal temperature.

Components Description of HRSG:

1. Steam Drums vs. Once Through Steam Generators (OTSG):

Steam Drums: It serves to separate steam from liquid water. Typically, there are

either two or three drums that produce steam at different pressures. Drum type

HRSG produce steam of approximately 5- 30% quality ion the HP evaporator or

boiler section. Steam is separated from liquid water by combined effects of

mechanical separators (cyclone or herring- bone panels) and gravity. Primary

steam separation is accomplished by centripetal deceleration of the steam/ water

mixture with the more dense water falling down through the cyclone and the steam

rising up the centre. The separated steam then passes through a secondary

separator, which serves the purpose of removing droplets that contain solids. Solids

are undesirable because they can cause fouling of downstream super-heater

headers, and fouling, erosion and corrosive deposits on downstream steam turbine

components.

Once Through Steam Generator (OTSG): An alternative to drum type HRSG is

the “once through steam generator (OTSG)”, in which feed water is converted

directly to super-heated steam without travelling through any drums.

Because there are no components with thick walls in an OTSG, the stresses from

thermal transients are reduced. Thus, the OTSG design is advantageous from cyclic

perspective. Another benefit is its ability to run dry for extended periods.

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2. Typical Flow Paths: The flow paths through an HRSG of both water/ steam mixtures and

the gas turbine exhaust will vary from plant to plant. They depend on such features as the

orientation of the exhaust gas path, the selection of steam drums vs. once through design,

the number of steam pressure levels, the presence of re-heaters, etc.

Water/ Steam Flow Path: Each HRSG receives feed water and heats it in a series

of boiler- tube assemblies to generate steam. To recover maximum amount of

thermal energy from the gas turbine exhaust, the HRSGs are designed for three

different steam pressure levels- high pressure (HP), intermediate pressure (IP) and

low pressure (LP). Each pressure level features its own super- heater section,

evaporator section and economiser section.

Flow path for the typical water/ steam cycle originates from the condensate system

and is pumped towards the HRSG, sometimes passing through a condensate pre-

heater or a de-aerator tank. Next, the water travels through the boiler feed pumps

and into the inlet of the HRSG economiser section.

There will be multiple tube assemblies in the economiser- perhaps four or five for

each of the three pressure levels- through which feed water passes in sequence.

From there, the water enters the respective drum (HP, IP or LP) through the feed

water inlet nozzle and continues onto the evaporator section.

Natural circulation is maintained in the evaporators by downcomers, which feed

the water from the drum through distribution manifolds down to the lower

evaporator headers. Steam is generated and naturally flows upward in the

evaporator tubes. The saturated steam/ water mixture is conducted from the upper

evaporator headers to the respective steam drums through risers. From there, the

saturated, dry steam passes through the respective super-heater section, where it is

heated above its saturation point to become super-heated. At this point, the flow

path changes for the three different sections:

HP super-heater steam passes through de-super heaters located between

super-heater sections, which are used to control the steam temperature, then

to the HP main steam line that feeds the HP steam turbine.

IP super-heater travels directly into the IP main steam line, where it is

combined with the “cold re-heat steam” coming from the exhaust of the HP

steam turbine.

LP super-heater steam travels directly into the LP main steam line that

feeds the LP steam turbine.

The cold re-heat steam is the steam that exits the discharge of the HP steam

turbine, after it has performed thermodynamic work. This cold re-heat

steam is combined with IP main steam, and the combined steam enters the

HRSG in the re-heater section. After the re-heater assembly, the steam is

directed to the IP and LP steam turbines to generate more power.

Gas side flow path: The flow path on the gas side is much simpler than the water/

steam side. The gas turbine exhaust enters the HRSG though the transition duct.

After passing through the transition duct, the exhaust gas passes across the various

pressure part sections of the HRSG. In doing so, it heats the water or steam inside

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the tubes by giving up its thermal energy to the working fluid. A typical unit have

the exhaust gas passing through the HRSG sections in the following order-

HP super-heater 1

Re-heater 1

Duct burner

HP super-heater 2

Re-heater 2

CO catalyst, if installed

HP evaporator

After passing across the HP evaporator tube assembly, the exhaust gas travels

through the ammonia injection grid, and then through the SCR system where the

mixture of exhaust gas and ammonia reacts with the NOx emissions. The exhaust

gas then passes across the remaining pressure part sections of the HRSG in the

following order-

LP super-heater

HP economiser 1

IP super-heater

HP economiser 2

IP evaporator

HP economiser 3

IP economiser 1

HP economiser 4

LP evaporator

LP economiser (or feed water pre-heater)

After the exhaust gas exits the final stage of economiser, it travels up

through the exhaust stack and it discharged to the atmosphere.

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Drum

Economiser

Evaporator

Re-heater

Superheater

Re-heater

Superheater

Cold reheat line

Spraywater

control

valve

HP Turbine

Condenser

Condensate

Pump

Boiler feed

pump

Feed water

tank

Feedwater

minimum

flow valve

economiser

economiser

Spraywater

control

valve

Feed water

control

valve

Drum outlet

valve

Attemperator

BOILER

Attemperator

STEAM/ WATER FLOW PATH

IP/LP

Turbine

Drum outlet

tank

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3. De-aerators: A de-aerating feed water heater is unusual in today’s large combined cycle

plants. Many plants are equipped with a de-aerating section in the main condenser.

Alternatively, the de-aerator may be integrating with the LP drum.

If the de-aeration is upstream of the boiler, the low feed water temperature results in de-

aeration performed under sub-atmospheric pressure. In this case, it is more practical to

vent the de-aerator into the condenser. The oxygen removal capabilities of this design are

excellent. However, carbon dioxide (CO2) which also enters the cycle with air ingress is

mostly condensable. Therefore, de-aeration under sub-atmospheric conditions is not very

efficient for removing CO2.

4. Economisers: The economiser (sometimes called the feed water pre-heater) extracts heat

from the exhaust gas stream just before it is discharged to the atmosphere and uses that

thermal energy to pre-heat the feed water. The benefits result: (1) the HRSG’s thermal

efficiency increases and (2) thermal stresses on the steam drum are reduced. The

economiser consists of modular finned tubes and header assemblies that are part of the

HRSG boiler and located within the casing. The number of tube rows per header and the

number of headers is determined by the design heating requirements of the unit.

Damage mechanisms:

Low Cycle Fatigue: The impingement of cold water on hot surfaces or vice-versa,

particularly during shut down and restart leads to this damage mechanism.

Differential Expansion: Uneven heating of tubes due to flow- or temperature

distribution problems can cause adjacent tubes to expand differently. Both

compressive and tensile loads are imposed.

Deposits: Poor water chemistry or excessive fast ramp rates can result in solid

precipitation and deposition, causing under heating.

Flow Accelerated corrosion: Single phase FAC in economisers is being

recognised as one cause of failure.

Corrosion Fatigue: Chemical imbalance can create corrosion and cyclic loading

can exacerbate the effect due to fatigue.

Erosion: Solids in the water can cause erosion at higher velocities.

5. Evaporators: Evaporators boil the water, turning it into steam at saturated conditions. In

an evaporator, heated water and steam flows up riser tubes and into the respective steam

drums (HP, IP or LP). The steam goes to the super-heater, while the water is re-

circulated via downcomers to the bottom of the evaporator.

The evaporator “pinch temperature” is what limits the amount of heat that can be

recovered in most HRSG designs. “Pinch Temperature” is defined as the difference

between the exhaust gas temperature leaving the evaporator and the steam saturation

temperature within the evaporator. The smaller the pinch temperature, the more efficient

the steam cycle, but also the higher the capital cost of the HRSG because of the

requirement of more heat transfer surface.

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Damage mechanisms:

Deposits: A major concern with evaporators is deposition in the inner tube walls,

most often on the upstream tube rows.

Low cycle fatigue: This occurs in natural circulation evaporators during start up

when sufficient piping flexibility is not provided to accommodate the temperature

differences that occur prior to circulation becoming fully established.

Differential expansion: The uneven heating of evaporator tubes- caused by

uneven distribution of (1) exhaust gas or steam/ water flows or (2) exhaust gas

temperatures- can cause adjacent tubes to expand or contract differently. Both

compressive and tensile loads are imposed.

Flow accelerated corrosion: Two phase flows in the evaporator cause FAC

particularly in low pressure section.

Erosion: Solids in the water and water in two phase flow systems can cause

erosion at higher velocities.

6. Super-heaters and Re-heaters: Purpose of the super-heater and re-heater sections of an

HRSG is to raise steam temperatures above saturation point in order to deliver maximum

energy to the steam turbine, and to eliminate moisture that could form in the steam as it

expands through the turbine, which would cause droplet impingement damage on steam

turbine components.

The super-heater and re-heater sections normally are located in the hottest gas stream, in

front of the evaporator. As a result, their tubes are exposed to the hottest exhaust gas and

thus experience the highest metal temperature. That is why super-heaters and re-heaters

require the most critical attention to material selection.

Damage mechanisms:

Thermal Fatigue: The impingement of hot gases on cold surfaces at start up or

of cold gases on hot surfaces at shut down creates thermal gradients. The high

pressure components are more vulnerable to fatigue effects due to their increased

wall thickness.

Thermal Shock: Thermal shock to the inner surfaces of the tubes and headers

can be caused by: condensate entering or remaining in a super-heater or re-heater

section; cold steam entering heat soaked, dry re-heater; and water from leaking or

malfunctioning de-superheaters.

Creep: Only high temperature components are prone to creep damage. Over

temperature transients and continuous high temperature operation will increase

the creep rate. However, if the creep is coupled with fatigue due to cycling, the

damage will be much higher.

Oxidation: Exposure of the metal to higher temperature than design designed

temperature result in oxidation. Oxidation and exfoliation can happen both inside

and outside the tubes and piping, caused by exhaust gas on one side and water/

steam on the other.

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Differential Expansion: The uneven heating of evaporator tubes- caused by

uneven distribution of (1) exhaust gas or steam/ water flows or (2) exhaust gas

temperatures- can cause adjacent tubes to expand or contract differently. Both

compressive and tensile loads are imposed.

Deposits: Drum carry over lead to the formation of deposits.

7. De-superheaters: De- superheating is the best way to control the outlet temperature of

an HRSG super-heater or re-heater. In doing so, se-superheater serve the vital purpose of

preventing thermal damage to super-heater and re-heater tubes as well as to outlet steam

piping and downstream equipment.

Precise control of spray water is essential to preventing equipment damage. It’s also

essential that spray water valves don’t leak by.

8. Steam By-pass Systems: During start up, shut down and steam turbine trips, the gas

turbine is producing so much exhaust heat at such rapid rate of temperature change that if

it were imposed uncontrolled onto the steam side, thermal ramp rate limits would be

violated. To resolve this imbalance, the HRSG is allowed to generate steam but then vent

it directly to the atmosphere until the steam side is properly warmed up.

In this scheme, the HP super-heated steam generated during start up is diverted around

the HP section. The steam temperature, through a pressure reducer and temperature

reducer and into the cold re-heats line where it cools the otherwise dry re-heater. The

primary job of the pressure reduction valve is to control the HP drum pressure, hence

limited thermal stresses on the HP drum.

In the second stage of bypass, the attemperated steam from cold re-heat line mixes with

the steam from the IP drum and is directed through the re-heater. After passing through

the re-heater, the steam travels through a second pressure control/ attemperating station

before it is directed through a dump tube into the condenser. This second pressure

control/ attemperator section is called hot re-heat bypass and its primary function is to

control cold re-heat pressure and prevent HP turbine windage heating.

A third bypass station diverts LP steam around the turbine and directly into the

condenser during plant transients.

Cascading bypass systems handle an immense amount of energy at elevated temperature,

pressure and velocities. They also experience harsh temperature and flow changes as

their pressure control valves open suddenly in response to plant transients. The result is

severe service for the valves and steam conditioning equipments.

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MECHANICAL DESIGN OF AN HRSG

ACCORDING TO

INDIAN BOILER REGULATIONS

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INDIAN BOILER REGULATIONS

History of Indian Boiler Regulations:

In the year 1863, a very serious boiler explosion occurred in Calcutta which caused the loss

of several lives. As a result of this explosion, the necessity of inspection of boilers was

widely recognised and a bill was introduced in the Bengal Council to provide for the

inspection of steam boilers. In the year 1864, the Bengal Act VI of 1864 was passed which

provided for the inspection of steam boilers and prime movers in the town and suburbs of

Calcutta. This is the beginning of boiler legislation in India.

Following the Bengal Act of 1864, each of the other provinces framed legislation. At that

time there were seven different Acts and seven different sets of rules and regulations. Those

Acts and rules & regulations were inconsistent with one another. As the differences in the

Acts and rules and regulations among the various provinces in India gave rise to many

difficulties and hampered the development of industries, the Central Government appointed a

committee called "The Boiler Law Committee" in 1920 to examine and report on the general

question of boiler legislation in India.

The Boiler Laws Committee, 1920-21, the first to review the boiler laws on a national scale

reported in March, 1921. The report criticised the differences in the Acts, rules and

regulations. The report also pointed out that in the inspection of boilers the personal element

was a weighty factor, and the difference in regulations resulted in what was termed as

"provincial jealousy". The report stressed that all provinces should be subject to the same

regulations and work done in one province should be accepted as correct in another province.

The Committee recommended that regulations to cover the standard conditions for material,

design and construction of boilers should be framed by Government of India and make

applicable to all the provinces. The report also pointed out that regulations were entirely of

technical nature and there was no reason for which these regulations would be affected by

local conditions. The Committee prepared a draft Act on the lines of which, the basic All-

India Act was passed in 1923. The Boiler Laws Committee also prepared a uniform set of

technical regulations and a model set of administrative rules. A sharp distinction was drawn

between the regulations and the rules. The regulations referred entirely to technical matters

where as the rules referred to questions concerning the administration of the Act. Indian

Boiler act, 1923 provides for the safety of life and property of persons from the danger of

explosion of boilers.

The Government of India Act, 1935 assigned the subject 'Boilers' to the concurrent field. The

provision for constituting Central Boilers Board having the authority to make regulations

consistent with the Act was made in the Indian Boilers (Amendment) Act, 1937. A Board

called the Central Boilers Board was accordingly constituted in the year 1937.

The Central Boilers Board in exercise of the powers conferred under section 28 of the said

Act, formulated regulations on boilers. The current version of these regulations is known as

the Indian Boiler Regulations, 1950 with amendments up to 22nd February, 2005.

21 | P a g e

INDIAN BOILER REGULATIONS

Regulations for Determining the Working Pressure to be Allowed on Various Parts

of Boilers Other Than Fusion Welded and Seamless Forged Drums:

Reg. No. 175:

Maximum pressure: The maximum pressure at which a boiler may be used shall be

determined in accordance with the provisions of this chapter. The regulations in this

chapter refer to material subjected to steam temperature not exceeding 500oF.

SHELLS

Reg. No. 176:

Formula for Working Pressure of Shells:

(a). For cylindrical shells, barrels, steam and water drums, and domes of boilers, the

maximum working pressure per square inch to be allowed shall be calculated from the

following formula:-

W. P. = (t - 2) x S x J

Eqn. (1)

C x D

Where

W. P. = the working pressure in lbs. per square inch;

t = the thickness of shell plates in 32nds of an inch;

S = the minimum tensile breaking strength of the shell plates in tons per square inch;

J = the percentage of strength of the longitudinal seams of shell or of a line of holes in the

shell for stays, or rivets, or of an opening in the shell not fully compensated, whichever is

least calculated by the methods hereafter described;

C is a co-efficient as follows:-

(1) 2.75 when the longitudinal seams are made with double butt straps and when small

shells are formed from solid rolled sections.

(2) 2.83 when the longitudinal seams are made with lap joints and are treble riveted.

(3) 2.9 when the longitudinal seams are made with lap joints and are double riveted

(4) 3.0 when the longitudinal seams are welded and are fitted with a single butt strap.

(5) 3.3 when the longitudinal seams are made with lap joints and are single riveted.

D = the inside diameter of the outer strake of plating of the cylindrical shell measured in

inches.

22 | P a g e

(b). The factor of Safety shall in no case be less than 4.

The actual factor of Safety in each case may be found from the equation ;

With the best form of joint and least co-efficient (c) the Factor of Safety for shell plates,

1/4 inch to 1-3/4 inches in thickness varies from 5.13 to 3.99.

Reg. No. 177:

Methods of Calculating the Strength of Riveted Joints:

(a). The percentage of strength of a riveted joint (J) shall be found from the following

formulae (i), (ii), (iii):

(i).

100(P - D)

= Plate percentage Eqn.(2).

P

(ii).

100 x A x N x C x S1

= Rivet percentage Eqn.(3)

P x T x S

(iii).

100(P -

2D) +

100 x A x C x S1 = Combined plate and rivet

percentage. Eqn.(4)

P P x T x S

Where

P is the pitch of rivets of outer row in inches.

D is the diameter of rivet holes in inches.

A is the sectional area of one rivet hole in square inches.

N is the number of rivets per pitch (P).

T is the thickness of plate in inches.

C = 1 for rivets in single shear as in lap joints, and 1.875 for rivets in double shear as in

F = 1.4 x C x

t

t - 2

23 | P a g e

double butt strapped joints.

S1 is the shearing strength of rivets which shall be taken to be 23 tons per square inch for

steel and 18 tons per square inch for iron.

S is the minimum tensile breaking strength of shell plates in tons per square inch.

In the first formula (i) D is the diameter of the rivet holes in the outer rows and in the third

formula D is the diameter of the rivet holes in the next rows. In the last formula A is the

area of one rivet hole in the outer row.

(b). When the sectional area of the rivet holes is not the same in all rows, and when some

of the rivets are in double shear and others in single shear the rivet sections per pitch of

each size in shear shall be computed separately and added together to form the total rivet

section.

Reg. No. 181:

Thickness of Butt Straps:

The minimum thickness of butt straps for the longitudinal seams of cylindrical shells shall

be determined by the following formulae but all straps should be of sufficient thickness to

permit efficient caulking and in any case shall not be less than 3/8 inch in thickness.

Single butt straps having ordinary riveting:-

1.125T = T1 Eqn.(5)

Single butt straps having every alternate rivet in the outer rows omitted:-

1.125T x (P – D)

= T1 Eqn.(6)

(P – 2D)

Double butt straps of equal width having ordinary riveting:-

0.625 T = T1 Eqn.(7)

Double butt straps of equal width having every alternate rivet in the outer rows omitted:-

0.625T x (P – D)

= T1 Eqn.(8)

(P – 2D)

Double butt straps of unequal width either having ordinary riveting, or having every

alternate riveting the outer rows omitted:-

0.75T = T1 (Wide strap) Eqn.(9)

0.625T = T1 (narrow strap) Eqn.(10)

24 | P a g e

Where

T1= the thickness of the butt straps in inches.

T= the thickness of plate in inches

P= the pitch of rivets at outer row in inches

D= the diameter of rivet holes in inches

Reg. No. 183:

Maximum Pitch of Rivets in longitudinal joints:

The maximum pith of the rivets in the longitudinal joints of boiler shells shall be:-

C x T + 1.625 = maximum pitch in inches Eqn. (11)

Where

T = the thickness of the shell plate in inches.

C is a co-efficient as given in the following table : -

Number of Rivets

Per pitch

Co -efficient for Lap

joints

Co-efficient for

single Butt- strapped

Joints

Co-efficient for

double Butt-

strapped joints

1

2

3

4

5

1.31

2.62

3.47

4.14

-

1.53

3.06

4.05

-

-

1.75

3.50

4.63

5.52

6.00

Reg. No. 185:

Circumferential and End Seams of Water Tube Boilers:

The suitability of circumferential seams including the seams joining ends to shells shall be

verified by the following formula : -

K x J x (t-2)

is equal to or greater than W.P. Eqn. (17)

D x C

Where

K = 150 for 26/30 tons tensile plates.

K = 157 for 28/32 tons tensile plates.

Due to higher stresses, see regulation 271 and 340

WP = the working pressure in lbs per sq. in.

25 | P a g e

D = the diameter of shell in inches, measured inside the outer ring of plates.

J = Circumferential Joint efficiency calculated by Eqn.2 or 3.

C = 8.24 where the seams are made with lap joints and are treble riveted.

= 8.44 where the seams are made with lap joints and are double riveted.

= 9.60 where the seams are made with lap joints and are single riveted.

t = thickness of plate in 32nd of an inch.

Reg. No. 186:

Compensation for Manholes and other openings:

The percentage of compensating section shall be determined by the following formulae:-

200(W - D) x Tr

= Percentage strength of compensating section Eqn. (18)

(L + 2D) x Ts

80 x A x N

= Percentage strength of rivet section Eqn. (19)

(L + 2D) x Ts

Where

W is the width of compensation ring in inches measured in the direction of the

longitudinal axis of the boiler,

L is the length of opening in shell in inches measured in the direction of the longitudinal

axis of the boiler,

D is the diameter of rivet holes in inches,

Tr is the thickness of compensation ring in inches,

Ts is the thickness of shell plate in inches.

A is the area of one rivet hole in inches,

N is the number of rivets on one side of the longitudinal line. When the rivets are in

double shear 1.875 times the single rivet section shall be allowed.

Parts of raised manhole mouthpieces within four inches of the shell shall, in addition to the

ring, be included in the compensating section.

26 | P a g e

Reg. No. 187:

Uncompensated holes in Water Tube Boilers:

The maximum diameter of any unreinforced opening shall not exceed 'd' subject to a

maximum of 203 millimetres.

The notations are as follows:-

K = P x D Eqn. 20

1.82 fe

where

P = working pressure;

d = maximum allowable diameter of opening (in the case of an

opening of elliptical or round form, the mean value of the

two axes of the opening shall be taken for d);

D = outer diameter of the shell;

e = actual thickness of the shell;

f = allowable stress;

When K has a value of unity or greater, the maximum size of an unreinforced opening

should be 51 millimetres (2 inches).

DISHED END PLATES

Reg. No. 188:

Complete hemisphere without stays or other support made of one or more plates and

subject to internal pressure:

The maximum working pressure shall be determined by the following formula : -

W. P. = (t - 2) x S x J Eqn.(23)

C x R

where

W.P. is the working pressure in lbs per square inch,

t is the thickness of the end plate in 32nds of an inch,

S is the minimum tensile breaking strength of the end plates in tons per

square inch, or whatever strength is allowed for them,

27 | P a g e

J is the least percentage of strength of the riveted joints of the plates

forming the hemisphere or securing it to be cylindrical shell

R is the inner radius of curvature in inches.

C for single riveting is 3.3,

C for double riveting is 2.9,

C for treble riveting is 2.83.

Reg. No. 189:

Dished Ends subject to Internal Pressure:

(a). For unstayed ends of steam and water drums, tops of vertical boilers, etc., when dished

to partial spherical form, the maximum working pressure shall be determined by the

following formula:-

W.P. = 15 x S x (t - 1) Eqn.(24)

R

Where

W.P. is the working pressure in lbs. per square inch,

t is the thickness of the end plate in 32nd of an inch,

S is the maximum tensile breaking strength of the end plates in tons per square inch,

R is the inner radius of curvature of the end in inches, which shall not exceed the external

diameter of the shell to which it is attached.

(b). The inside radius of curvature at the flange shall not be less than 4 times the thickness

of the end plate, and in no case less than 2½ inches.

(c). When the end has a manhole in it, (t-5) shall be substituted for (t-1) in the formula.

(d). The total depth of flange of manhole from the outer surface in inches measured on the

minor axis shall be at least equal to:-

(T x W) 1/2 = depth of flange in inches. Eqn.(25)

Where

T is the thickness of the plate in inches, and

W is the minor axis of the hole in inches.

28 | P a g e

Reg. No. 214:

Efficiency of Ligaments:

When a shell or drum is drilled for tubes in a line parallel to the axis of the shell or drum,

the efficiency of the ligament between the tube holes shall be determined as follows:-

(a). When the pitch of the tube holes on every row is equal, the formula is:-

p - d

= efficiency of ligament Eqn.(54)

p

Where

p= pitch of tube holes in inches

d= diameter of tube holes in inches.

The pitch of tube holes shall be measured either on the flat plate before rolling or on the

median line after rolling.

Example:- Pitch of tube holes in the drums as shown in fig.= 5 ¼” , diameter of tube= 3

¼” , diameter of tube holes=3 9/32”.

p - d

=

5.25 - 3.281

= 0.375, efficiency of ligament.

p 5.25

5 ¼” 5 ¼” 5 ¼” 5 ¼” 5 ¼” 5 ¼” 5 ¼”

Longitudinal line

29 | P a g e

(b). When the pitch of tube holes on any one row is unequal, the formula is:-

p -nd

= efficiency of ligament Eqn.(55)

p

Where

p= unit length of ligament in inches,

n= number of tube holes in length p,

d= diameter of tube holes in inches.

Example: - spacing show in fig. Diameter of tube holes= 3 9/32”

p - nd 12

=

12/ 2 x 3.281

= 0.453, efficiency of ligament.

p 12

Example:- Spacing shown in fig. Diameter of tube hole = 3-9/32".

p – nd

=

29.25 - 5 x 3.281

= = 0.439, efficiency of ligament

P 29.25

5 ¼” 6 ¼” 5 ¼” 6 ¼” 5 ¼” 6 ¼” 5 ¼”

12”

Longitudinal line

30 | P a g e

(c). The strength of those ligaments between the tube holes subjected to a longitudinal

stress shall be at least one-half the required strength of those ligaments which come

between the two holes which are subject to circumferential stress.

5 ¼”

29 ¼”

Longitudinal line

5¼

” 6 ¼” 5¼

” 6 ¼” 5¼

” 6 ¼” 5¼

”

5¼

”

31 | P a g e

Fusion Welded And Seamless Forged Drums For Water Tube Boilers And Super

Heaters:

SHELLS

Reg. No. 270:

Shell of Steam and Water Drum:

(a). The working pressure shall be determined by the following formula:-

W. P. = 2 f E (t - .03) Equation (72)

D + T - .03

Where

T= Thickness in inches.

D = Maximum internal diameter in inches.

W. P.= Working pressure in lb. per sq.inch.

f = Permissible working stress in lbs. per sq. inch at working metal temperature.

E = The efficiency of ligaments between the tube holes or other uncompensated openings

in shell, or the weld factor of the longitudinal joints.

For Class I boilers, the weld factor shall be taken as 1

In the particular case of an un-pierced wrapper plate of a fusion welded drum.

E = 1

Reg. No. 271:

Permissible working stresses for shells of Boiler and Integral Super-heater Drums

and Headers:

The maximum permissible stress for drum shells and headers shall be taken as available in

governing Boiler codes of the country of the material to which it belongs. In case of non-

availability of the value, the following procedure for evaluating shall be adopted.

(i). For temperatures at or below 454o C, the smaller of the two values:-

Et R

f = ------ or ------

1.5 2.7

(ii). For temperatures above 454o C, the least of the following three values:-

Et SR

(a) ----- (b) ----- and (c) Sc

1.5 1.5

where,

t = Working metal temperature,

R = Minimum specified tensile strength of the steel at room temperature

E = Minimum specified Yield point at room temperature

Et = Yield Point (0.2% proof stress) at the temperature’t’.

32 | P a g e

Sc = The average stress to produce an elongation of 1% (creep) in 100,000 hours at

temperature 't'.

SR = The average stress to produce rupture in 100,000 hours at the temperature 't' and in

no case more than 1.33 times the lowest stress to produce rupture at the temperature.

The working metal temperature shall be taken as:-

(a). For saturated steam, water and mud drums, the saturation temperature corresponding

to the pressure WP plus 50oF.

(b). For superheated steam the designed maximum steam temperature for that drum plus

50oF.

STANDPIPES AND NOZZLES:

Reg. No. 279:

Standpipes and Nozzles welded to shell:

Where the standpipes and nozzles are secured by welding, adequate compensation for the

hole cut shall be provided. Compensation shall be considered adequate when the sectional

area ‘X’ to be compensated measured through the axis of the shell is less than the

compensating area ‘Y’ according to figure.

33 | P a g e

Sectional area X = ( dn x es )

Sectional area Y = 2 ( tn - ea ) b x

fn +

fn + 2 ( ta - es ) C1 ÷ C3

fs fs

Where,

dn = internal diameter of the standpipe or nozzle;

ts = actual thickness of shell;

tn = actual thickness of standpipe or nozzle;

es = equivalent thickness of shell i.e., thickness of a seamless shell of similar material un-

pierced by tube holes and is designed for the same conditions of pressure and temperature

as the shell in question;

en = equivalent thickness of the standpipe calculated similarly as for e plus any thickness if

required to withstand any external load on the standpipe or nozzle;

b = the least of the value of

2.5 times t ;

2.5 times tn ;

or if the length of the standpipe or nozzle outside or inside the shell is less than this

value, ‘b’ shall be limited to the actual length in each case.

Cw = the aggregate cross sectional area of the fillet welds.

dn

Ci = ts + 76 mm (3in.) or -----, whichever is greater;

2

fs = the permissible stress for the material of the shell at the working metal temperature.

fn = the permissible stress for the material of the standpipe or nozzle at the working metal

temperature.

In cases where ‘Y’ is less than ‘X’ a compensating plate shall be fitted to the shell at the

standpipe and secured by fillet welds as in figure.

The area of cross section of this compensation plate shall be governed by the ratio between

the permissible stress at the working metal temperature for the material of the shell and the

compensating plate.

34 | P a g e

Boiler and Super-heater tubes, Headers and other Pressure Part tubes:

Reg. No. 338:

The working pressure of the tubes shall be determined by the following formula:-

W. P. = 2 f (T - C) Eqn.(87)

(D - T + C)

Where

T = minimum thickness of tubes, that is, nominal thickness less than the permissible

negative tolerance in mm. (inch)

C = 0.75 mm for working pressure up to and including 70 kg/cm².

or

C = 0 for working pressure exceeding 70 kg/cm² (1000 lbs/sq. in.)

W. P.= Working pressure of boiler in kg/cm² (lbs/sq. inch)

D = External diameter of tube in mm (inch)

f= permissible stress for the material at the working metal temperature in kg/cm² (lbs/sq.

Inch) to be determined on the basis given below:-

The working metal temperature shall be taken as :-

(a) For integral economiser tubes, the maximum water temperature for which the part of

the element is designed plus 11°C (20°F).

(b) For furnace and boiler tubes, the saturation temperature corresponding to the working

pressure plus 28°C (50o F)

(c) For convection superheater tubes, the maximum steam temperature for which the part

of the element is designed plus 39oC (70oF).

(d) For radiant superheater tubes the designed maximum steam temperature plus 50 °C

(90oF)

Permissible working stress for tubes:- For temperatures at or below 454 °C,

T. S.

or

Et whichever is lower.

2.7 1.5

For temperature above 454 °C,

Sr

or Sc whichever is lower.

1.5

where,

T.S. = Minimum tensile strength of the material at room temperature.

Et = Yield point (0.2% proof stress) at working metal temperature 't'.

35 | P a g e

Sr = the average stress to produce rupture in 100,000 hours and in no case more than

1.33 times the lowest stress to produce rupture at the working metal temperature.

Sc = the average stress to produce an elongation of 1% (creep) in 100,000 hours, at the

working metal temp.

Note:- In case Sc values are not available in Material Standard and such materials are

known to have been used in boilers in India or abroad, then for such material the

allowable stress may be taken as the lower of

Et

or

Sr

1.5 1.5

Reg. No. 340:

Rectangular headers symmetrical in form:

The working pressure shall not exceed the smaller of the value obtained from the

following formula:-

(i) WP = c f (t - c1 )² Eqn. 88(a)

b²

(ii) WP = 6.25 t² f E Eqn. 88(b)

W²

Where,

W. P. = working pressure.

t = thickness.

b = internal breadth between the supporting sides of the header.

c = 3.413 for wrought steel and 2.926 for steel castings.

Where the sides are corrugated or otherwise reinforced by substantial supports so that the

length of the portion between the corrugations or supports does not exceed b, shall be

taken as 5.82 for wrought steel and 5.12 for steel castings.

f = permissible stress at working metal temperature.

c1 = 0.08cm.(0.03in.)

W = internal width of the header between the tube plate and the opposite side.

E = efficiency of ligaments between the holes.

36 | P a g e

End Attachments.-The ends of rectangular headers may be formed integral with the

header or attached by welding.

The working pressure for the flat ends shall be calculated by the following formula:-

W. P. = f ( t - C )² Eqn.(89A)

d² K

Where

WP = Working Pressure.

f = Permissible stress for the material at the working metal temperature.

t = thickness of the plate at the weakest part.

d = the least width between the walls of the rectangular header.

C = 1 mm (0.04”).

K = 0.32 for ends integral with or flanged and butt welded to the header.

= 0.40 for ends directly strength welded to the header in an approved manner.

Reg. No.341

Headers Irregular in Form:

In cases where the headers are of such irregular form as to render impracticable the

application of a formula for the determination of thickness, the manufacturer shall show

the suitability of the headers for the working conditions by indicating practically the

maximum internal hydraulic pressure which a header, made to the same design and of

similar material, will withstand without permanent deformation. The maximum

working pressure for similar headers may then be determined by the following

formula:-

W. P. = P1 x

Permissible stress at working metal

temperature Eqn.(90)

F x C

where,

W. P. = The working pressure in lbs. Per sq. In.

P1 = The maximum internal hydraulic pressure withstood without permanent

deformation.

F = 1.75 for wrought steel and 2 for cast steel.

C = 15500 for wrought steel of 24 tons per sq. in minimum ultimate tensile stress.

= 16500 for wrought steel of 26 tons per sq. in minimum ultimate tensile stress.

= 18000 for wrought steel of 28 tons per sq. in minimum ultimate tensile stress.

= 21000 for wrought steel of 32 tons per sq. in minimum ultimate tensile stress.

= 22000 for wrought steel of 34 tons per sq. in minimum ultimate tensile stress.

= 15500 for wrought steel of 28 tons per sq. in minimum ultimate tensile stress

37 | P a g e

.

MECHANICAL DESIGN OF AN HRSG

ACCORDING TO

“AMERICAN SOCIETY

OF

MECHANICAL ENGINEERS”

(ASME SEC I & SEC VIII)

38 | P a g e

Introduction to ASME:

The ASME code design criteria consist of basic rules specifying the design method, design

loads, allowable stress, acceptable materials, fabrication, testing, and certification and

inspection requirements.

The design method known as “design by rule” uses design pressure, allowable stress and a

design formula compatible with the geometry to calculate the minimum required thickness of

pressurised thanks, vessels and pipes.

The ASME- American Society of Mechanical Engineers- International Boiler and Pressure

Vessel Code is made of 12 sections and contains over 15 divisions and subsections.

Code Sections:

1. Power Boilers

2. Materials

3. Rules for Construction of Nuclear Facility Components

4. Heating Boilers

5. Non- destructive Examination

6. Recommended Rules for the Care and Operation of Heating Boilers

7. Recommended Guidelines for the Care of Power Boilers

8. Pressure Vessels

9. Welding and Brazing Qualification

10. Fibre-reinforced Plastic Pressure Vessels

11. Rules for In-service Inspection of Nuclear Power Plant Components

12. Rules for Construction and Continued Service of Transport Tanks

39 | P a g e

ASME Sec I and Sec VIII- Fundamentals:

The formulae in ASME Section I and Section VIII are used to determine the minimum

required thickness and design pressure of piping, tubes, drums and headers using Maximum

Allowable Working Pressure (MAWP). These formulae may also be used for calculating

wall thickness of tubes and pipes under internal pressure.

Design:

The ASME Boiler code section I and Section VIII requires longitudinal and circumferential

butt joints to be examined by full radiography.

When the vessel design is required fully radiographed longitudinal butt-welded joint, the

cylindrical shell will have a joint efficiency factor (E= 1.0). This factor corresponds to a

safety factor of 3.5 in the parent material.

When the vessel design is required non- radiographed longitudinal butt- welded joint, the

vessel will have a joint efficiency factor (E= 0.7), which corresponds to a safety factor of

0.5 resulting in an increase of 43% in the thickness of the plates.

Pressure Vessels Maximum Allowable Stress Values:

The maximum allowable stress values to be used in the calculation of the vessel’s wall

thickness are given in the ASME Code for many different materials. These stress values are a

function of temperature.

Division 1: governs the design by Rules, is less stringent from the standpoint of certain

design details and inspection procedures, and thus incorporates a higher safety factor of 4.

For example, if a 60,000 psi tensile strength material is used, the Maximum Allowable

Stress Value is 15,000 psi.

Division 2: governs the design by Analysis and incorporates a lower safety factor of 3.

Thus, the maximum allowable stress value for a 60,000 psi tensile strength material will

become 20.000 psi.

40 | P a g e

Maximum Allowable Stress Value for Common Steels

Material

Spec. Nbr

Grade

DIVISION 1

-20oF to

650oF

DIVISION 2

-20oF to

650oF

Carbon Steel

Plates and

Sheets

SA- 516

Grade 55 13,800 18,300

Grade 60 15,000 20,000

Grade 65 16,300 21,700

Grade 70 17,500 23,300

SA- 285

Grade A 11,300 15,000

Grade B 12,500 16,700

Grade C 13,800 18,300

SA- 36

12,700

16,900

SA-203

Grade A 16,300 21,700

Grade B 17,500 23,300

Grade D 16,300 21,700

Grade E 17,500 23,300

High Allow

Steel Plates

SA- 240

Grade 304 11,200 20,000

Grade 304L - 16,700

Grade 316 12,300 20,000

Grade 316L 10,200 16,700

ASME Sec I- Power Boilers: Types, Design Fabrication,

Inspection and Repair:

ASME Sec I- Boiler Tubes up to and including 5 inches O.D. (125 mm):

a) The minimum required thickness use equation below:

t =PD

2S + P+ 0.005D + e 𝐞𝐪. 𝟏

b) To calculate the Maximum Allowable Working Pressure (MAWP):

P = S [2t−0.01D−2e

D−(t−0.005D−e)] 𝐞𝐪. 𝟐

41 | P a g e

Where,

t= Minimum Design Wall Thickness (in)

P= Design Pressure (psi)

D= Tube Outside Diameter (in)

e= Thickness Factor (0.04 for expanding tubes; 0 for strength welded tubes)

S= Maximum Allowable Stress

ASME Sec I- Piping, Drums, and Headers:

a) Using the outside diameter:

t =PD

2SE + 2yP+ C 𝐞𝐪. 𝟑

P =2SE(t − C)

D − (2y)(t − C) 𝐞𝐪. 𝟒

b) Using the inside radius:

t =PR

SE − (1 − y)P+ C 𝐞𝐪. 𝟓

P =SE(t − C)

R + (1 − y)(t − C) 𝐞𝐪. 𝟔

Where,

t= Minimum design wall thickness (in)

P= Design pressure (psi)

D= Tube outside diameter (in)

R= Tube radius (in)

E= Tube welding factor (1.0 for seamless pipe; 0.85 for welded pipe)

y= Wall thickness welding factor (0.4 for 900oF & lower; 0.7 for 950oF & up)

C= Corrosion allowance (0 for no corrosion; 0.0625 in., commonly used; 0.125 in.,

maximum

S= Maximum Allowable Stress.

42 | P a g e

ASME Sec VIII- Division I, Division 2, Division 3

The ASME Sec VIII, rules for fired or unfired pressure vessels, is divided into three

divisions to provide the requirements applicable to the design, fabrication, inspection, testing

and certification. The following formulae and allowable stresses are only for Division I, the

main code.

Sec VIII- Thin Cylindrical Shells:

The formulae in ASME Sec VIII, Division I, used for calculating the wall thickness and

design pressure of pressure vessels are:

a) Circumferential Stress (longitudinal welds):

When P < 0.385SE:

t =PR

(SE − 0.6P) 𝐞𝐪. 𝟕

P =SEt

(R + 0.6t) 𝐞𝐪. 𝟖

(R= Internal Radius)

b) Longitudinal Stress (circumferential welds):

When P < 1.25SE:

t =PR

(2SE + 0.4P) 𝐞𝐪. 𝟗

P =2SEt

(R − 0.4t) 𝐞𝐪. 𝟏𝟎

(R= Internal Radius)

43 | P a g e

Sec VIII- Thick Cylindrical Shells:

For internal pressure higher than 3,000 psi, special considerations are specified. As the

ratio of t/R increases beyond 0.5, an accurate equation is required to determine the

thickness.

The formulae used for calculating thickness of wall and design pressure are:

a) For longitudinal welds:

When P > 0.385SE

t = R { Z12 − 1} where Z =

(SE + P)

(SE − P) 𝐞𝐪. 𝟏𝟏

And

P = SE {(Z − 1)

(Z + 1)} where Z = [

(R + t)

R]𝟐 𝐞𝐪. 𝟏𝟐

b) For circumferential welds:

When P > 1.25SE

t = R {Z12 − 1} where Z = (

P

SE) + 1 𝐞𝐪. 𝟏𝟑

And

P = SE (Z − 1) where Z = [(R + t)

R]

𝟐

𝐞𝐪. 𝟏𝟒

44 | P a g e

ASME Sec I- Pressure Piping- Minimum Wall Thickness:

According to ASME Sec I, the minimum thickness of piping under pressure is

t =PD

2SE + 2yP+ C 𝐞𝐪. 𝟏𝟓

Where

t (min)= Minimum wall thickness required (in)

P= Design Pressure (psig)

D= Outside diameter of Pipe (in)

S= Allowable Stress in pipe material (psi)

E= Longitudinal joint factor- E=1.0 for seamless pipe, E= 0.85 for ERW pipe

C= Corrosion allowance, typically 0.05 in

y= Wall thickness co-efficient

= 0.4 for T ≤ 900oF

= 0.5 for 900 < T ≤ 950oF

= 0.7 for 950 < T ≤ 1000oF

ASME Sec VIII- Reinforcement Wall Thickness Plate:

The standard design method uses an increased wall thickness plate at the equator line of the

vessel to support the additional stresses caused by the attachment of the legs. The formula for

the calculation of the wall thickness of a segmented plate to be welded in a vessel or spherical

shell is:

45 | P a g e

t =PL

2SE − 0.2P+ C 𝐞𝐪. 𝟏𝟔

L = Di/2

Where,

t= Minimum Design Wall Thickness (in)

P= Design Pressure (psi)

Di= Inside Diameter of Sphere (in)

L= Sphere Radius

E= Tube Welding Factor (1.0 for seamless pipe; 0.85 for welded pipe)

C= Corrosion Allowance (0- no corrosion; 0.0625 in. commonly used; 0.125 in. maximum)

S= Maximum Allowable Stress

ASME Sec I- Dished Head Formulae:

Flanged and dished heads can be formed in a size range from 4 in. to 300 in. in diameter and

in thickness range of 14 gauge to 1 ½” thick. Pressure vessel heads and dished ends are

essentially the same- the end caps of a pressure vessel tank or an industrial boiler, supplied

with a flanged edge to make it easier for the fabricator to weld the head to the main body of

the tank.

Dished heads can be manufactured using a combination of processes, spinning, and flanging,

where the spherical radius is made via the spinning process and the knuckle is created under

the flanging method.

Blank, Unstayed Dished Heads:

The thickness of a blank, unstayed dished head with the pressure on the concave side, when it

is a segment of a sphere, shall be calculated by the following formula:

t =5PL

4.8S 𝐞𝐪. 𝟏𝟕

Where,

t= Minimum thickness of head (in)

P= Maximum allowable working pressure (psi)

L= Concave side radius (in)

S= Maximum allowable working stress (psi)

46 | P a g e

Seamless or Full-Hemispherical Head:

The thickness of a blank, unstayed, full-hemispherical head with the pressure on the concave

side shall be calculated by the formula:

t =PL

2S − 0.2P 𝐞𝐪. 𝟏𝟖

Where,

t= Minimum thickness of head (in)

P= Maximum Allowable Working Pressure (psi)

L= Radius to which the head was formed (in)

S= Maximum Allowable Working Stress (psi)

Note: the above formula shall not be used when the required thickness of the head given by

the formula exceeds 35.6% of the inside radius. Instead, use the following formula:

t = L (Y13 − 1) where Y =

2(S + P)

2S − P 𝐞𝐪. 𝟏𝟗

ASME Sec VIII- Division 1: Dished Head Formulae:

The ASME Sec VII- Division 1 determines the rules for dished heads. The most common

configurations are spherical, hemispherical, elliptical (or ellipsoidal) and torispherical shapes.

47 | P a g e

Spherical or Hemispherical Heads:

When t < 0.356R or P < 0.665SE (Thin Spherical or Hemispherical Heads):

t =PR

2SE − 0.2P 𝐞𝐪. 𝟐𝟎

And

P =2SEt

R + 0.2t 𝐞𝐪. 𝟐𝟏

When t > 0.356R or P > 0.665SE (Thick Spherical or Hemispherical Heads):

t = R (Y13 − 1) where Y =

2(SE + P)

2SE − P 𝐞𝐪. 𝟐𝟐

And

P = 2SE (Y − 1

Y + 2) where Y = (

R + t

R)

3

𝐞𝐪. 𝟐𝟑

Elliptical or Ellipsoidal Heads- Semi-Elliptical or Semi- Ellipsoidal Heads-2:1:

The commonly used semi-ellipsoidal head has a ratio of base radius to depth of 2:1. The

actual shape can be approximated by a spherical radius of 0.9D and a knuckle radius of

0.17D. The required thickness of 2:1 heads with pressure on the concave side is given below:

48 | P a g e

t =PD

2SE − 0.2P 𝐞𝐪. 𝟐𝟒

And

P =2SEt

D + 0.2r 𝐞𝐪. 𝟐𝟓

Torispherical Heads:

Shallow heads, commonly referred to as flanged and dished heads (F&D heads), are with a

spherical radius ‘L’ of 0.1D and a knuckle radius ‘r’ of 0.06D.

Flanged & Dished Head (F&D heads):

The dished radius of a flanged and dished head is 0.1D and the knuckle radius is 0.06D.

the required thickness of a torispherical F&D head with r/L= 0.06 and L=Di, is

t =0.885PL

SE − 0.1P 𝐞𝐪. 𝟐𝟔

And

P =SEt

0.885L + 0.1t 𝐞𝐪. 𝟐𝟕

Where,

P= Pressure in the concave side of the head

S= allowable stress

t= Thickness of the head

L= Inside Spherical Radius

E= Joint Efficiency Factor

Non-Standard 80-10 Flanged and Dished Head:

49 | P a g e

On an 80-10, the inside radius (L) is 0.8Di and the knuckle radius (ri) is 10% of the

head diameter. For the required thickness of a non standard 80-10 head, use equation 26

and 27.

Conical or Toriconical Heads:

The required thickness of the conical or toriconical head (knuckle radius > 6% OD) shall be

determined by formula using internal diameter of shell, α≤ 300.

t =PD

2(SE − 0.6P)cosα 𝐞𝐪. 𝟐𝟖

L= Di / (2cosα)

Di = Internal Diamter (conical portion)= D- 2r (1- cosα)

r= Inside knuckle radius

50 | P a g e

ASME Sec VIII- Shell Nozzles:

Vessel components are weakened when material is removed to provide openings for nozzles

or access openings. To avoid failure in the opening area, compensation or reinforcement is

required. The code procedure is to relocate the removed material to an area within an

effective boundary around the opening. Figure below shows the steps necessary to reinforce

an opening in a pressure vessel.

Definitions:

51 | P a g e

Diameter of Circular Opening, d:

d= Diameter of Opening – 2 (Tn + Corrosion Allowance)

Required Wall Thickness of the Nozzle (min.):

tn =PR

SE − 0.6P

Area of Required Reinforcement, Ar:

Ar = d. ts. F

Where,

d= Diameter of circular opening, or finished dimension of opening in plane under

consideration (in.)

ts= Minimum required thickness of shell when E= 1.0 (in.)

F= correction factor, normally 1.0

52 | P a g e

DESIGN CALCULATIONS

53 | P a g e

Example 1-Boiler Tube:

Calculate the minimum required wall thickness of a water tube boiler 2.75 in. O.D., strength

welded (E or e=0) into place in a boiler. The tube has an average wall temperature of 650oF.

The maximum allowable working pressure (MAWP) is 580 psi gauge. Material is carbon

steel SA-192.

Solution:

For tubing up to and including 5.in. O.D., use equation 1.

P=580 psi

D=2.75in.

e= 0 (strength welded)

S=11,800 psi at 650oF

Thus,

t =PD

2S + P+ 0.005D + e

t =2.75 × 580

2(11,800) + 580+ 0.005(2.75) + 0

t = 0.079 in.

Example 2- Steam Piping:

Calculate the required minimum thickness of a seamless steam piping at a pressure of 900

psi gauge and a temperature of 700oF. The piping is 10.77 in. O.D. plain end; the material is

SA-335-P1, allow steel. Allow a manufacturer’s tolerance allowance of 12.5%.

Solution:

P= 900 psi

D= 10.77 in.

C= 0

S= 13,800 psi at 700oF

E= 1.0

y= 0.4 (ferric steel less than 900oF)

Thus, using equation 3,

t =PD

2SE + 2yP+ C

54 | P a g e

t =900 × 10.77

2(13,800)(1.0) + 2(0.4)(900)+ 0

t= 0.34 in.

Therefore, including a manufacturer’s allowance of 12.5%; - 0.34 X 1.125 = 0.38 in.

Example 3-Maximum Allowable Working Pressure (MAWP):

Calculate the Maximum Allowable Working Pressure (MAWP) for a seamless steel pipe of

material SA- 209-T1. The outside pipe diameter is 12.75 in. with a wall thickness of 0.46

in. The operating temperature is 850oF. The pipe is plain ended.

Solution:

D= 12.75 in.

t= 0.46 in.

C= 0 (3 to 4 inches nominal and larger)

S= 13,100 psi at 850oF

E= 1.0 (seamless pipe)

y= 0.4 (austenitic steel at 850oF)

Using equation 4,

P =2SE(t − C)

D − (2y)(t − C)

P ={2(13,100)(1.0) × (0.46 − 0)}

12.75 − (2 × 0.4)(0.46 − 0)

P =26,200 × 0.46

12.75 − 0.368

P= 973 psi

55 | P a g e

Example 4- Welded Tube Boiler Drum:

A welded water tube boiler drum of SA-516-60 material is fabricated to an inside radius of

18.70 in. on the tube side and 19.68 in. on the drum. The plate thickness of the tube sheet

and drum are 2.34 in. and 1.49 in. respectively.

The longitudinal joint efficiency is 100% and the ligament efficiencies are 56% horizontal

and 30% circumferential. The operating temperature is not to exceed 500oF. Determine the

Maximum Allowable Working Pressure (MAWP) based on:

Welded Water Tube Boiler Drum: DRUM & TUBE SHEET

Solution:

This example has two parts:

a) The Drum- consider the drum to be plain with no penetrations.

b) The Tube Sheet- consider the drum to have penetrations for boiler tubes.

a) Drum. Use equation 6 (inside radius R)

P =SE(t − C)

R + (1 − y)(t − C)

Where

S=15,000 psi at 500oF

E=1.0

t=1.49 in.

C= 0

R=19.68 in.

y= 0.4 (ferric steel less than 900oF)

Drum P ={(15,000)(1.0)(1.49 − 0)}

19.68 + (1 = 0.4)(1.49 − 0)

Drum P =15,000 × 1.49

19.68 + 0.894

Drum P= 1,068 psi.

56 | P a g e

b) Tube Sheet. Use equation 6 (inside radius R)

P =SE(t − C)

R + (1 − y)(t − C)

Where,

S=15,000 psi

E=0.56 (circumferential stress=30% and longitudinal stress=56%; therefore 0.56< 2 X 0.30)

t= 2.34 in.

C= 0 (3 to 4 inches nominal and larger)

R= 18.70 in.

y= 0.4 (ferric steel less than 900oF)

Tubesheet P =(15,000)(1.0)(2.34 − 0)

18.70 + (1 − 0.4)(2.34 − 0)

Tubesheet P =15,000 × 2.34

18.70 + 1.404

Tubesheet P = 1.746 psi

NOTE: Consider the Maximum Allowable Working Pressure (MAWP)= 1,086 psi

(lowest number).

Example 5- Thin Cylindrical Shell:

A vertical boiler is constructed of SA-515-60. It has an internal diameter of 96 in. and an

internal design pressure of 1000 psi at 450oF. The corrosion allowance is 0.125 in. and joint

efficiency is E= 0.85. Calculate the required thickness of the shell.

Solution:

Consider S = 15,000 psi

Since P < 0.385SE as 1000 psi < 6,545 psi, use equation 7:

t =PR

(SE − 0.6P)+ C

Considering the internal radius = 48 in. and corrosion allowance = 0.125 in.

t =1000 × 48

2(15000)(0.85) − 0.6(1000)+ 0.125

t = 2.052 in.

57 | P a g e

Example 6-Thick Cylindrical Shell:

Calculate the required thickness of an accumulator with P= 10,000 psi, R =18 in. , S=

20,000 psi and E = 1.0. Assume corrosion allowance of 0.125 in.

Solution:

For P > 0.385SE as 10,000 psi > 7,700 psi, use equation 11

t = R { Z12 − 1} where Z =

(SE + P)

(SE − P)

Z =(20,000)(1.0) + 10,000

(20,000)(1.0) − 10,000=

30,000

10,000= 3

Therefore,

t = (18) (312 − 1) + 0.125

t = 8.08 in.

Example 7-Thick Cylindrical Shell:

Calculate the required thickness of an accumulator with P= 7,650 psi, R =18 in. , S= 20,000

psi and E = 1.0. Assume corrosion allowance of 0 in.

Solution:

For P < 0.385SE as 7,650 psi < 7,700 psi, use equation 7

t =PR

(SE − 0.6P)+ C

t =7,650 × 18

(20,000)(1.0) − (0.6)(7,650)+ 0

t = 8.9 in.

58 | P a g e

Example 8- comparison between eq. 7 and eq. 11:

Calculate the shell thickness from the data in the previous example, comparing the equation 7

with another answer using equation 11.

Solution:

t = R { Z12 − 1} where Z =

(SE + P)

(SE − P)

Z =(20,000)(1.0) + 7,650

(20,000)(1.0) − 7,650=

27,650

12,350= 2.24

Therefore,

t = (18 + 0) (2.2412 − 1)

t = 8.9 in.

This shows that the simple use of equation 7 is accurate over a wide range of R/t ratios.

Example 9-Segment of a Spherical Dished Head:

Calculate the thickness of a seamless, blank, unstayed dished head having pressure on the

concave side. The head has an inside diameter of42.7 in. with a dish radius of 36.0 in. The

maximum Allowable Working Pressure (MAWP) is 360 psi and the material is SA-285 A.

The temperature does not exceed 480oF. State if this thickness meets code.

Solution:

Using equation 17-

t =5PL

4.8S

P= 360 psi

L= 36.0 in.

S= 11,300 psi at 480oF

t =5(360 × 36)

4.8(11,300)

𝐭 = 𝟏. 𝟏𝟗 𝐢𝐧.

59 | P a g e

Note: No head, except a full- hemisphere head, shall be of lesser thickness than that

required for a seamless pipe of the same diameter. Then, the minimum thickness of piping

is:

t =PD

2SE + 2yP+ C

y= 0.4 (ferric steel < 900oF; E=1.0)

t =360 × 42.7

2(11,300)(1.0) + 2(0.4)(360)

t = 0.67 in.

Therefore, the calculated head thickness meets the code requirements, since 1.19 > 0.67.

Example 10-Segment of a Spherical Dished Head with a Flanged- in Manhole:

Calculate the thickness of a seamless, unstayed dished head with pressure on the concave

side, having a flanged- in manhole 6.0 in. X 16 in. Diameter head is 47.5 in. with a dish

radius of 45 in. The MAWP is 225 psi, the material is SA-285-C, and temperature does not

exceed 428oF.

Solution:

First thing to check is the radius of the dish is at least 80% of the diameter of the shell:

Dish Radius

Shell Diameter=

45

47.5

60 | P a g e

0.947 > 0.8- the radius of the dish meets the criteria. Then using equation 17:

t =5PL

4.8S

P= 225 psi

L= 45 in.

S= 13,800 psi at 450oF (use the next higher temperature)

t =5(225 × 45)

4.8(13,800)

𝐭 = 𝟎. 𝟕𝟔𝟒 𝐢𝐧.

This thickness is for a blank head. According to ASME this thickness is to be increased by

15% or 0.125 in. whichever is greater. As 0.764 X 0.15 = 0.114 in. that is less than 0.125

in., then, the thickness must be increased by 0.125 in.

Thus, the required head thickness is t = 0.764 + 0.125 = ~ 0.90 in.

Example 11-Seamless or Full- Hemispherical Head:

Calculate the minimum required thickness for a blank, unstayed full- hemispherical head. The

radius to which the head is dished is 7.5 in. The MAWP is 900 psi and the head material is

SA-285-C. The average temperature of the header is 570oF.

Solution:

Using equation 18-

t =PL

2S − 0.2P

P= 900 psi

L=7.5 in.

S=13,800 at 600oF

t =900 × 7.5

2(13,800) − 0.2(900)= 𝟎. 𝟐𝟒 𝐢𝐧.

Check, if this thickness exceeds 35.6% of the inside radius = 7.5 X 0.356 = 2.67 in. It does

not exceed 35.6%; therefore, the calculated thickness of the head meets the code

requirements.

61 | P a g e

Example 12-Thin Spherical or Hemispherical Head:

A pressure vessel is built of SA- 516-70 material and has an inside diameter of 96 in. The

internal design pressure is 100 psi at 450oF. Corrosion allowance is 0.125 in. and joint

efficiency is E = 0.85. Calculate the required spherical head thickness of the pressure vessel

if ‘S’ is 20,000 psi.

Solution:

Since 0.665SE > P = 11305 psi > 100 psi, use equation 20-

The inside radius in a corroded condition is equal to, R = 48 + 0.125 = 48.125 in.

t =PR

2SE − 0.2P

t =100 × 48.125

2(20,000)(0.85) − 0.2(100)=

𝐭 = 𝟎. 𝟏𝟒 𝐢𝐧.

Example 12.1-Thin Spherical or Hemispherical Head:

A spherical pressure vessel with an internal diameter of 120 in. has a head thickness of 1 in.

Determine the design pressure if the allowable stress is 16,300 psi. Assume joint efficiency

E= 0.85. No corrosion allowance is stated to the design pressure.

Solution:

Since t < 0.356R, using equation 21-

P =2SEt

R + 0.2t

P =2(16,300)(0.85)(1)

60 + 0.2(1)

𝐏 = 𝟒𝟔𝟎 𝐩𝐬𝐢

The calculated pressure 460 psi < 0.665SE (9213 psi); therefore, equation 21 is acceptable.

62 | P a g e

Example 13-Thick Hemispherical Head:

Calculate the required hemispherical head thickness of an accumulator with P= 10,000

psi, R= 18.0 in., S= 15,000 psi and E= 1.0. Corrosion allowance is 0.125 in.

Solution:

As P (10,000 psi) > 0.665SE (9975 psi), using equation 22,

t = R (Y13 − 1) where Y =

2(SE + P)

2SE − P

Y =2[(15,000)(1.0) + 10,000]

2[(15,000)(1.0)] − 10,000=

50,000

20,000= 2.5

Therefore,

𝐭 = 𝐑 (𝐘𝟏𝟑 − 𝟏) = 18.0 (2.5

13 − 1) = 𝟔. 𝟓𝟓𝟒 𝐢𝐧.

Example 14-Elliptical or Ellipsoidal Heads, or Semi-Elliptical or Semi- Ellipsoidal

Heads:

Calculate a semi-elliptical head thickness considering the dimensions given below:

Inside Diameter of Head (Di): 18.0 in.

Inside Crown Radius (L): (18.0 X 0.9Di) in.

Inside Knuckle Radius (ri): (18.0 X 0.17Di) in.

Straight Skirt Length (h): 1.500 in.

Radius L- (18.0 X 0.9Di) = 16.20 in.

Radius ri- (18.0 X 0.17Di) = 3.06 in.

63 | P a g e

Material and Conditions:

Material: SA-202 Gr. B (room temperature)

Internal Pressure: 200 psi

Allowable Stress: 20,000 psi

Head Longitudinal Joint Efficiency: 0.85

Corrosion Allowance: 0.010 in.

Solution:

Variable:

L/r = L/ ri = 16.20/3.06 = 5.29 in.

Now,

Required Thickness (using equation 24)

t =PD

2SE − 0.2P+ corrosion allowance

t =200 × 18.0

2(20,000)(0.85) − 0.2(200)+ 0.010

t= 0.116 in.

Maximum Pressure (using equation 25)

P =2SEt

D + 0.2r

P =2(20,000)(0.85)(0.116)

18 + 0.2(0.116)

𝐏 = 𝟐𝟏𝟗 𝐩𝐬𝐢

Example 15- Torispherical Head:

A drum is to operate at 500oF and 350 psi and to hold 5000 gallons of water. The inside

radius of the Dished Torispherical Head is 78 in. The material is SA 285 Grade A.

Assume S= 11,200 psi and E = 0.85.

Solution:

Dished Torispherical Head with L = Di and r/L = 0.06. Using equation 26,

t =0.885PL

SE − 0.1P=

0.885(350)(78)

(11,200)(0.85) − 0.1(350)= 𝟐. 𝟓𝟒 𝐢𝐧.

64 | P a g e

Example 16-Basic Pipe Nozzle:

Basic Design:

Design Pressure: 300 psig

Design Temperature: 200oF

Shell Material is SA-516 Gr. 60

Nozzle Diameter: 8 in. Sch. 40

Nozzle Material is SA-53 Gr. B Seamless

Corrosion Allowance= 0.0625”

Vessel is 100% Radiographed

Solution:

a) Wall thickness of the nozzle (min.)

tn =PR

SE − 0.6P

tn =300 × 4.312

12,000(1.0) − 0.6(300)+ 0.0625 (corrosion allowance)

tn = 0.11 + 0.0625 = 0.17 in. (min) – Pipe Sch. 40 is t = 0.32 in.

65 | P a g e

b) Circular opening, d:

d= Diameter of Opening – 2 (Tn + Corrosion Allowance)

d= 8.625 – 2 (0.32 + 0.0625)

d= 8.625 – 2(0.3825)

d= 7.86 in.

c) Area of required reinforcement, Ar:

Ar = d. ts. F

Ar = 7.86 × 0.487 × 1.0 = 𝟑. 𝟖𝟐 𝐢𝐧𝟐

Available reinforcement area in shell, Ar, as larger of As or An:

𝐀s = 𝐋𝐚𝐫𝐠𝐞𝐫 𝐨𝐟: 𝐝(𝐓𝐬 − 𝐭𝐬) − 𝟐𝐓𝐧(𝐓𝐬 − 𝐭𝐬) 𝒐𝒓 𝟐(𝐓𝐬 + 𝐓𝐧)(𝐓𝐬 − 𝐭𝐬) − 𝟐𝐭𝐧(𝐓𝐬 − 𝐭𝐬)

As = 7.86 (0.5625 – 0.487) – 2 X 0.5625 (0.5625 – 0.487) = 0.50 in2

𝐀𝐧 = 𝐒𝐦𝐚𝐥𝐥𝐞𝐫 𝐨𝐟: 𝟐[𝟐. 𝟓(𝐓𝐬)(𝐓𝐧 − 𝐭𝐧)] 𝐨𝐫 𝟐[𝟐. 𝟓(𝐓𝐧)(𝐓𝐧 − 𝐭𝐧)]

An = 2 [2.5 (0.5626) (0.32 – 0.17)] = 0.42 in2

Ar < (As + An)

Here,

As + An = 0.50 + 0.42 = 0.92 in2

Thus, Ar > (As+ An)

Therefore, it is necessary to increase Ts and/ or Tn to attend the premise Ar < (As + An)

66 | P a g e

Example 17- Basic Shell and Nozzle:

Design Pressure = 700 psi

Design Temperature = 700oF

Nozzle Diameter = 8 in. (8.625 OD)

Material:

Shell- SA 516 Gr. 70

Head- SA 516 Gr. 70

Nozzle- SA 106 Gr. B

E = 1.0 (weld efficiency)

Required Shell Thickness:

𝐭𝐬 =𝐏𝐑

𝐒𝐄 − 𝟎. 𝟔𝐏

ts =700 × 30

16,600(1.0) − 0.6(700)= 𝟏. 𝟑𝟎 𝐢𝐧.

Required Head Thickness:

𝐭𝐡 =𝐏𝐑

𝟐𝐒𝐄 − 𝟎. 𝟐𝐏

th =700 × 30

2(16,600)(1.0) − 0.2(700)= 𝟎. 𝟔𝟒 𝐢𝐧.

Required Nozzle Thickness:

𝐭𝐧 =𝐏𝐑

𝐒𝐄 − 𝟎. 𝟔𝐏

tn =700 × 4.312

14,400(1.0) − 0.6(700)= 𝟎. 𝟐𝟏 𝐢𝐧.

Opening Reinforcement:

Ar = d. ts = 8.625 X 1.3 = 11.2 in2

As = d (Ts – ts) – 2 Tn (Ts – ts)

As = 8.625 (1.5 – 1.3) – 2 (1.0) (1.5 – 1.3)

= 1.325 in2

An = 2[2.5 (Ts) (Tn – tn)]

An = 2[2.5 (1.5) (1.0 – 0.21)] = 5.925 in2

Ar < (As + An)

Here,

As+ An = 1.325 + 5.925 = 7.25 in2

Thus, Ar > (As + An)

Therefore, it is necessary to increase Ts and/ or

Tn to attend to premise Ar < (As + An )

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PROJECT INFERENCE

Thus, a Heat Recovery Steam Generator (HRSG) has been successfully designed according

to the Indian Boiler Regulations (IBR) and the American Society of Mechanical Engineers

(ASME Section I and Section VIII).

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CONCLUSION

The internship for three months at Thermax Limited, Pune has been a very eye-opening one

as it has provided me a peek into the practical working conditions of the industry. It has

provided me with the chance to have a firsthand experience of how an industry functions

and also to learn about the practical applications of my field of study. This will surely go a

long way in moulding me as a successful engineer in the future.

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REFERENCE

Indian Boiler Regulation, 1950

ASME Section I and Section VIII

www.google.com

www.wikipedia.org

User’s handbook for Heat Recovery Steam Generator (HRSG)