Ghinwa JASSAR-Jonny Saidy.pdf

8/16/2019 Ghinwa JASSAR-Jonny Saidy.pdf

1/81

Lebanese University

Faculty of Engineering II

Final year project

Submitted in fulfillment of the requirements for the

Mechanical Engineering Degree

By

Ghinwa JASSAR

Johnny SAIDY

Design of an exhaust gas heat recovery fire tube boiler for

Poppins® factory

Project supervisor: Dr. Khalil EL KHOURY

2012


2/81


3/81

i

Acknowledgments

Foremost, we would like to thank Dr. Khalil Khoury, chief of the mechanical depart-

ment, for helping us to learn how to think and research. He has been a mentor and an

inspiration. His encouragement and support made this work possible. We especially appreci-

ate the project opportunity that he has given us as well as his faith in our abilities.

We are also hearty thankful and express deep sense of gratitude to all the Daher

Group managers and staff for the project opportunity and for their moral support and inter-

est, particularly to Mr. Albert Sassine, engineering manager, for his invaluable guidance and

technical advice. Without his ideas, the project would not have appeared in the present

shape.

We would also like to thank our fellow graduate students for their support and

friendship.

We would like to thank our family, who have continually given us their love and sup-

port, and encouraged us to reach our dreams. We could not have done this without you.

Most importantly we would like to thank God. Thank you for all of these blessings.


4/81

ii

Abstract

The mid-20th century has proven to be a time during which the world has had a rude awak-

ening from its relaxed attitude towards the usage of our depleting natural resources. Proofof this is the waste heat recovery systems that have been in use in industries all over the

world for the last 50 years. An example of this is the integration of various factory sections

where the waste heat from one section is used in another.

Moreover, basic human needs can be met only through industrial growth, which depends to

a great extent on energy supply. The large increase in population during the last few decades

and the spurt in industrial growth have placed tremendous burden on the electrical utility

industry and process plants producing chemicals, fertilizers, petrochemicals, and other es-

sential commodities, resulting in the need for additional capacity in the areas of power and

steam generation throughout the world. Steam is used in nearly every industry, and it is wellknown that steam generators and heat recovery boilers are vital to power and process

plants. It is no wonder that with rising fuel and energy costs engineers in these fields are

working on innovative methods to generate electricity, improve energy utilization in these

plants, and recover energy efficiently from various waste gas sources

The study of improved heat transfer performance is referred to as heat transfer enhance-

ment . In general, this means an increase in heat transfer coefficient. Attempts to increase

heat transfer coefficients have been recorded for more than a century, and there is a large

store of information. A survey [1] cites 4345 technical publications. The recent growth of

activity in this area is clearly evident from the yearly distribution of the publications present-ed in Figure 1.


5/81

iii

References on heat transfer augmentation versus year of publication (to late 1990) [1].

Waste heat recovery is common practice in the food industry and not only saves money, but

streamlines production and results in better efficiencies.

The definition of waste heat includes the following:

1. Unburned combustible fuel.

2. Sensible and latent enthalpy discharge from exhaust gas mixtures.

3. Sensible heat discharge in liquid waste.


6/81

iv

Nomenclature

Coefficient of thermal expansion

A Heat exchanger surface areaC p Specific heat capacityd Tube diameterD Shell diameter

ΔP Pressure drop through the boiler

ε Absolute roughness of tubesЄ Heat exchanger effectivenessE Joint efficiency between sheets (section 4.3)E Modulus of elasticity (section 7.2)

f Darcy friction factor

h Water and Steam enthalpyh Thickness of the tube sheet (section 7.2)k Exhaust gas thermal conductivityK Steel thermal conductivityK Minor losses coefficient (section 4.3.2)L Tube lengtḣ Mass flow rate Ligament efficiency n Number of tubesNu Nusselt numberPr Prandtl number

Q Boiler dutyr Heat loss factor through the outer shell ρ Exhaust gas densityRe Reynolds numberR f Fouling resistance

S Maximum allowable stress of steel (According to ASME code)t Shell thicknessU Overall heat transfer coefficientV Gas velocity inside the tubes

ν Kinematic Viscosity of exhaust gas

Differential thermal expansion

ubscripts

c Cold fluidh Hot fluid

i Inside fluido Outside fluidm Mean value

e Equivalent

p Tube sheet ( Ep)

s Shell (Es)t Tube (Et )


7/81

v


8/81

vi

Table of contents

1 INTRODUCTION ....................................................................................................................................... 1

2 OVERVIEW ON HEAT EXCHANGERS ......................................................................................................... 2

2.1 INTRODUCTION.............................................................................................................................................. 2

2.2 HEAT EXCHANGER CLASSIFICATION ............................................................. ....................................................... 2

2.3 HEAT EXCHANGERS DESIGN METHODS .............................................................................................................. 3

2.3.1 Overall Heat Transfer Coefficient .................................................................................................... 3

2.3.2 Convective Heat Transfer Coefficient h ........................................................................................... 4

2.3.3 LMTD Method.................................................................................................................................. 5

2.3.4 Heat Exchanger Pressure Drop ........................................................................................................ 6

2.3.5 Analysis of Extended Surfaces ......................................................................................................... 6

2.3.6 Fouling in heat exchangers .............................................................................................................. 6

2.3.7 Typical Heat Exchanger Designs ...................................................................................................... 7

2.3.8 Waste Heat Recovery Boilers ........................................................................................................ 13

3 PROJECT DATA ...................................................................................................................................... 15

3.1 PROJECT DESCRIPTION .................................................................................................................................. 15

3.2 INPUT DATA ............................................................................................................................................... 19

3.2.1 Exhaust Gas ................................................................................................................................... 19

3.2.2 Feedwater...................................................................................................................................... 22

3.3 REQUIRED OUTPUT DATA .............................................................................................................................. 23

3.4 PRELIMINARY CALCULATION ........................................................... ................................................................ 24

4 HEAT TRANSFER DESIGN ....................................................................................................................... 27

4.1 FIRE TUBE BOILER SIZING PROCEDURE .............................................................................................................. 27

4.2 NUMERICAL APPLICATIONS ............................................................................................................................ 31

4.2.1 Gas and steam properties ............................................................................................................. 31

4.2.2 Mass flow per tube ........................................................................................................................ 31


9/81

vii

4.2.3 Exhaust Gas Velocity ..................................................................................................................... 32

4.2.4 The Reynolds Number.................................................................................................................... 32

4.2.5 The Nusselt Number and Dittus-Boelter Correlation ..................................................................... 33

4.2.6 Gas side heat transfer coefficient h i .............................................................................................. 33

4.2.7 Water side heat transfer coefficient ho ......................................................................................... 33

4.2.8 Overall heat transfer coefficient U ................................................................................................ 36

4.2.9 Log mean temperature difference ΔT LM ........................................................................................ 37

4.2.10 Boiler Duty Q ............................................................................................................................. 37

4.2.11 Surface Area and Tube Length .................................................................................................. 37

4.3 PRESSURE DROP .......................................................................................................................................... 38

4.3.1 Darcy Friction Factor ..................................................................................................................... 39

4.3.2 Minor Losses and Equivalent Length Le ......................................................................................... 41

4.3.3 Pressure Drop ................................................................................................................................ 43

4.3.4 Heat Transfer vs. Pressure Drop .................................................................................................... 44

5 OFF-DESIGN PERFORMANCE ................................................................................................................. 47

6 BOILER’S MECHANICAL DESIGN ............................................................................................................. 52

6.1 PRESSURE VESSEL ......................................................................................................................................... 52

6.1.1 Introduction ................................................................................................................................... 52

6.1.2 Manufacturing Constraints ........................................................................................................... 52

6.1.3 Shell Dimensions ............................................................................................................................ 52

6.1.4 Material and Maximum Allowable Stress ..................................................................................... 53

6.1.5 Loadings and Design Pressure ....................................................................................................... 54

6.1.6 Pressure Vessel thickness under internal pressure ........................................................................ 55

6.2 TUBE SHEET DESIGN ..................................................................................................................................... 56

6.2.1 Design Procedure........................................................................................................................... 57

6.2.2 Working Conditions ....................................................................................................................... 58

6.2.3 Numerical Application ................................................................................................................... 59


10/81

viii

7 BOILER FITTINGS AND MOUNTINGS ...................................................................................................... 62

7.1 SAFETY VALVES ............................................................................................................................................ 62

7.2 BOILER STOP VALVES..................................................................................................................................... 63

7.3 FEEDWATER CHECK VALVE ............................................................. ................................................................ 63

7.4 PRESSURE GAUGE ........................................................................................................................................ 64

7.5 WATER LEVEL INDICATOR .............................................................................................................................. 65

7.6 WATER LEVEL CONTROLS ............................................................................................................................... 65

7.7 AIR VENTS AND VACUUM BREAKERS ................................................................................................................. 65

8 CONCLUSION ......................................................................................................................................... 67

REFERENCES ................................................................................................................................................... 69


11/81

ix

List of Figures

Figure 2-1 Double pipe heat exchanger ..................................................................................... 8

Figure 2-2 Single-pass Shell and Tube heat exchanger .............................................................. 9

Figure 2-3 Finned tube heat exchanger ..................................................................................... 9

Figure 2-4 A boiler is basically a burner and a heat exchanger ............................................... 10

Figure 2-5 One pass fire tube boiler ......................................................................................... 11

Figure 2-6 Two pass fire tube boiler ........................................................................................ 11

Figure 2-7 Natural water circulation in a water tube boiler .................................................... 12

Figure 2-8 Water tube boiler schematic .................................................................................. 12

Figure 2-9 Comparison between fire tube and water tube boilers ......................................... 13

Figure 2-10 classification of waste heat boilers ....................................................................... 14

Figure 3-1 Caterpillar Diesel Engine ......................................................................................... 15

Figure 3-2 Diesel Engine Specifications .................................................................................... 16

Figure 3-3 Boiler placement ..................................................................................................... 17

Figure 3-4 The boiler without the outer shell .......................................................................... 18

Figure 3-5 Close-up on the fire tube boiler .............................................................................. 18

Figure 3-6 Diesel Engine Technical Data .................................................................................. 19

Figure 4-1 Fire tube boiler design procedure flow chart ......................................................... 30

Figure 4-2 Nusselt, Grashof and Prandtl numbers ................................................................... 34

Figure 4-3 Natural convection heat transfer from an isothermal horizontal cylinder ............ 34

Figure 4-4 Excel spreadsheet calculations for natural convection .......................................... 35

Figure 4-5 Entrance flow conditions and loss coefficient (a) Reentrant, K=0.8, (b) sharp-

edged, K=0.5, (c) slightly rounded, K=0.2, (d) well-rounded, K=0.04. ..................................... 42

Figure 4-6 Exit flow conditions and loss coefficient (a) Reentrant, K=1, (b) sharp-edged, K=1,

(c) slightly rounded, K=1, (d) well-rounded, K=1. .................................................................... 42

Figure 4-7 Influence of various geometrical parameters of a shell-and-tube exchanger on

heat transfer and pressure drop. ............................................................................................. 45

Figure 4-8 Influence of various geometrical parameters of a shell-and-tube exchanger on

heat transfer and pressure drop. ............................................................................................. 46

Figure 5-1 Flow chart for fire tube design procedure .............................................................. 48

Figure 6-1 Tubesheet AutoCad drawing and dimensions ........................................................ 53

Figure 6-2 types of stresses in a cylindrical shell, S1=longitudinal stress, S2=Circumferential

or hoop stress ........................................................................................................................... 55

Figure 6-3 Fire tube boiler scheme .......................................................................................... 57

http://d/FYP/RRRRRRREEEEEEEEEEPPPPPPPPPPOORT.docx%23_Toc332892214http://d/FYP/RRRRRRREEEEEEEEEEPPPPPPPPPPOORT.docx%23_Toc332892214http://d/FYP/RRRRRRREEEEEEEEEEPPPPPPPPPPOORT.docx%23_Toc332892214http://d/FYP/RRRRRRREEEEEEEEEEPPPPPPPPPPOORT.docx%23_Toc332892222http://d/FYP/RRRRRRREEEEEEEEEEPPPPPPPPPPOORT.docx%23_Toc332892222http://d/FYP/RRRRRRREEEEEEEEEEPPPPPPPPPPOORT.docx%23_Toc332892222http://d/FYP/RRRRRRREEEEEEEEEEPPPPPPPPPPOORT.docx%23_Toc332892222http://d/FYP/RRRRRRREEEEEEEEEEPPPPPPPPPPOORT.docx%23_Toc332892214


12/81

x

Figure 7-1 Boiler safety valve ................................................................................................... 62

Figure 7-2 Boiler stop valve ...................................................................................................... 63

Figure 7-3 Boiler Check Valve ................................................................................................... 63

Figure 7-4 Location of feed check valve ................................................................................... 64

Figure 7-5 Typical pressure gauge with ring siphon ................................................................ 64

Figure 7-6 Gauge glass and fittings .......................................................................................... 65

Figure 7-7 Typical air vents and vacuum breakers ................................................................... 66

List of Tables

Table 1 - Shape Factors .............................................................................................................. 4

Table 2 - TEMA Design Fouling Resistances Rf for a Number of Industrial Fluids ..................... 7

Table 3 Exhaust gas major and minor constituents ................................................................. 20

Table 4 Density, specific heat, thermal conductivity, expansion coefficient, kinematic

viscosity and Prandtl of exhaust gas (N2=76%; CO2=13%; H2O=11%) .................................... 21

Table 5 - Compressed water property table at 0.6 MPa.......................................................... 22

Table 6 - Saturated steam pressure table ................................................................................ 24

Table 7 - Results of design calculations for fire tube waste heat boilers for the same duty .. 44

Table 8 Values for G1, G2, G3 and G4 ...................................................................................... 58


13/81

1

1 Introduction

In our project, we’ve considered working with “Daher International Food Co.”, a leading

producer and distributor of premium quality food in the region, and particularly one of its

brands: Poppins®.

The Poppins® factory lies in the heart of the Bekaa valley in Mansoura, West Bekaa-Lebanon.

Poppins® uses the latest equipment and methods in the production of a wide array of break-

fast cereals and cereal chocolate bars.

However, due to chronic electricity shortage in Lebanon, frequent power outages occur on

daily basis outside of the country's capital Beirut (3 to 4 times per day).

These frequent power outages have a very undesirable impact on the factory’s production

line and cause the production chain to stop for some time as well as the production of non-

completed products and putting many delicate and relatively expensive systems in risk of

failure. All this imposes meaningful expenses and loss of productivity in the factory.

Hence, factories find themselves obligated to generate their own power using Diesel Engines

Generators in order to provide an uninterrupted power supply that has become costly now-

adays due to the global increase in fuel cost.

In fact, it is well-known that approximately one third of the total energy released by thecombustion process is lost along with the exhaust gas, and one third is transferred to the

cooling fluids while the rest is converted to actual electrical power. The cooling fluids are

necessary losses which prevent catastrophic failure of the engine due to overheating.

In our detailed study, the sensible heat must be recovered from the hot exhaust gas (around

500˚C) by means of a heat exchanger which will impose a low back pressure on the exhaust

system in order to prevent motor failure. We note that latent enthalpy recovery due to the

condensation of vapor is not envisaged at present in our project.

The final system design would ideally convert the recovered energy into heat energy that

is capable of producing fair amounts of pressurized steam that will be used in the cookingprocess of cornflakes. The waste heat recovery heat exchanger will be connected to the

main steam boilers network present in the factory.


14/81


15/81


16/81

4

where the surface efficiency of inner and outer surfaces, h is the heat transfer coefficientsfor the inner and outer surfaces, and S is a shape factor for the wall separating the two flu-

ids.

The surface efficiency accounts for the effects of any extended surface which is present on

either side of the parting wall.The thermal resistances include: the inner and outer film resistances, inner and outer ex-

tended surface efficiencies, and conduction through a dividing wall which keeps the two

fluid streams from mixing. The shape factors for a number of useful wall configurations are

given below in Table 1.

This equation is for clean or unfouled heat exchanger surfaces. The effects of fouling on heat

exchanger performance are discussed in a later section. Finally, we should note that:

However,

Table 1 - Shape Factors

Finally, the order of magnitude of the thermal resistances in the definition of the overall

heat transfer coefficient can have a significant influence on the calculation of the overall

heat transfer coefficient. Depending upon the nature of the fluids, one or more resistances

may dominate making additional resistances unimportant.

2.3.2 Convective Heat Transfer Coefficient h

The heat transfer coefficient, in thermodynamics and in mechanical and chemical engineer-

ing, is used in calculating the heat transfer, typically by convection. The heat transfer

coefficient has SI units in watts per square meter -kelvin: W/ (m2K).


17/81


18/81

6

Where and represent the temperature difference at each end of the heat exchang-er, whether parallel flow or counterflow. The LMTD expression assumes that the overall heat

transfer coefficient is constant along the entire flow length of the heat exchanger.

The LMTD method is also applicable to crossflow arrangements when used with the cross-

flow correction factor.

2.3.4 Heat Exchanger Pressure Drop

Pressure drop in heat exchangers is an important consideration during the design stage.

Since fluid circulation requires some form of pump or fan, additional costs are incurred as a

result of poor design.

In addition, as it is the case in our study, high pressure drop in the heat exchanger can

cause a high backpressure on the Diesel generators causing them to shut down.

Pressure drop calculations are required for both fluid streams, and in most cases flow con-

sists of either two internal streams or an internal and external stream. Pressure drop is

affected by a number of factors, namely the type of flow (laminar or turbulent) and the pas-

sage geometry.

First, a fluid experiences an entrance loss as it enters the heat exchanger core due to a sud-

den reduction in flow area, then the core itself contributes a loss due to friction and other

internal losses, and finally as the fluid exits the core it experiences a loss due to a sudden

expansion. In addition, if the density changes through the core as a result of heating or cool-

ing an acceleration or deceleration in flow is experienced.

This also contributes to the overall pressure drop (or gain). All of these effects are discussed

in detail later on in section 4.3.

2.3.5 Analysis of Extended Surfaces

Extended surfaces also known as fins are widely used as a means of decreasing the thermal

resistance of a system. In the study of heat transfer, a fin is a surface that extends from an

object to increase the rate of heat transfer to or from the environment by increasing convec-

tion. The amount of conduction, convection, or radiation of an object determines the

amount of heat it transfers. Increasing the temperature difference between the object and

the environment, increasing the convection heat transfer coefficient, or increasing the sur-

face area of the object increases the heat transfer. Sometimes it is not economical or it is not

feasible to change the first two options. Adding a fin to an object, however, increases the

surface area and can sometimes be an economical solution to heat transfer problems.

2.3.6 Fouling in heat exchangers

Fouling in heat exchangers represents a major source of performance degradation. Fouling

not only contributes to a decrease in thermal efficiency, but also hydraulic efficiency. The

buildup of scale or other deposit increases the overall thermal resistance of the heat ex-changer core which directly reduces the overall thermal efficiency. If buildup of a fouling


19/81

7

deposit is significant, it can also increase pressure drop due to the reduced flow area in the

heat exchanger core. The two effects combined can lead to serious performance degrada-

tion. In some cases the degradation in hydraulic performance is greater than the

degradation in thermal performance which necessitates cleaning of the heat exchanger on a

regular basis. Fouling of heat exchangers has different aspects. The two most common are

corrosion and scale buildup. However, depending upon the nature of the fluid other factorsmay contribute to fouling. Fouling in heat exchangers is traditionally treated using the con-

cept of a fouling resistance. This resistance is added in series to either side of the wall

resistance in the definition of the overall heat transfer coefficient.

Where and are respectively the inside and outside fouling resistances of the heatexchanger’s surface area.

Some typical values of fouling resistances are given in Table 2 for a number of fluids.

Table 2 - TEMA Design Fouling Resistances Rf for a Number of Industrial Fluids

2.3.7 Typical Heat Exchanger Designs

We will now examine several common heat exchanger designs without any detailed study.

We shall consider: Double Pipe Heat Exchangers, Shell and Tube Heat Exchangers, Compact

Heat Exchangers, Plate and Frame Heat Exchangers, Boilers, Condensers, and Evaporators.


20/81

8

2.3.7.1 Double Pipe Exchangers

The double pipe heat exchanger is probably one of the simplest configurations found in ap-

plications. It consists of two concentric circular tubes with one fluid flowing inside the inner

tube and the other fluid flowing inside the annular space between the tubes. Its primary

uses are in cooling process fluids where small heat transfer areas are required. It may bedesigned in a number of arrangements such as parallel flow and counterflow, and combined

in series or parallel arrangements with other heat exchangers to form a system.

Figure 2-1 Double pipe heat exchanger

2.3.7.2 Shell and Tube Exchangers

Shell and tube heat exchangers are widely used as power condensers, oil coolers, preheat-

ers, and steam generators. They consist of many tubes mounted parallel to each other in a

cylindrical shell. Flow may be parallel, counter, or cross flow and in some cases combinations

of these flow arrangements as a result of baffling. Shell and tube designs are relatively sim-

ple and most often designed according to the Tubular Exchanger Manufacturer’s Association

(TEMA) standards. Special attention must be given to the internal tube arrangement, i.e.

baffled, single pass, multi-pass, tube pitch and arrangement, etc., to properly predict theheat transfer coefficient.


21/81

9

Figure 2-2 Single-pass Shell and Tube heat exchanger

2.3.7.3 Finned Tube Heat Exchangers

In a conventional tube-fin exchanger, heat transfer between the two fluids takes place by

conduction through the tube wall.

In a gas-to-liquid exchanger, the heat transfer coefficient on the liquid side is generally one

order of magnitude higher than that on the gas side. Hence, to have balanced thermal con-

ductance on both sides for a minimum-size heat exchanger, fins are used on the gas side to

increase surface area A. This is similar to the case of a condensing or evaporating fluid

stream on one side and gas on the other. In addition, if the pressure is high for one fluid, it isgenerally economical to employ tubes.

Figure 2-3 Finned tube heat exchanger


22/81

10

2.3.7.4 Boilers

Basically, a boiler is a closed vessel or arrangement of enclosed tubes in which water is heated to

supply steam (to drive an engine or turbine, or to provide heat); when other liquid than water is

used, the boiler is more often named vaporizer (or evaporator). A second meaning of boiler is a do-

mestic device burning solid fuel, gas, or oil, to provide hot water, especially for central heating(better called a heater). Closely related to boilers are pressure cookers, i.e. strong hermetically

sealed pots in which food may be cooked quickly under pressure at a temperature above the normal

boiling point of water (in this case the intention is not to supply steam but to generate it for pressur-

izing; the higher the pressure, the higher the boiling temperature).

Most boilers are fuel-fired, thus, they can be viewed as shell-and-tube heat exchangers (Fig-

ure 2-5.), where the hot fluid is the burnt gases, and the cold fluid the water stream. Heat

transfer by radiation is important in boilers because of the high temperatures (some 2000 K).

In most boilers, the air for combustion is previously heated by the exhaust gases in the stack.

Typical efficiencies, measured as water enthalpy change divided by the combustion enthalpy

(most often based on the standard low heating value of the fuel), are around 100% in mod-

ern condensation boilers (where part of the water vapor dissolved in the flue gases is

condensed), around 90% for large non-condensing boilers, and around 80% for modern

small non-condensing boilers. A boiler is often the largest energy consumer both at domestic

and at industrial level, thus, great savings may be obtained by their proper selection, opera-

tion and maintenance.

Figure 2-4 A boiler is basically a burner and a heat exchanger

Two main types of boilers will be discussed in details:

Fire Tube Boilers

Water Tube Boilers

Fire Tube Boilers

In a fire-tube boiler (Figure 2-5, 2-6), hot flue gases from the burner are channeled through

tubes that are surrounded by the fluid to be heated. The body of the boiler is the pressure

vessel and contains the fluid. In most cases this fluid is water that will be circulated for heat-

ing purposes or converted to steam for process use. Fire-tube boilers are relatively

inexpensive, easy to clean, and more compact than water-tube boilers (although of smaller

steam capacities, and not suitable for high pressure applications (up to 2 MPa only).


23/81

11

Figure 2-5 One pass fire tube boiler

Figure 2-6 Two pass fire tube boiler

Water tube boilers

In a water-tube boiler, water flows through the tubes within a furnace in which the burner

fires into. The tubes are connected to a steam drum on top and a mud drum at the bottom.

Water-tube boilers typically produce steam or hot water for large industrial applications

(less frequently for heating applications).

Many water-tube boilers operate on the principle of natural water circulation (also known as'thermosiphoning'). Figure 2-7 helps to explain this principle:


24/81

12

-

Figure 2-7 Natural water circulation in a water tube boiler

Cooler feed water is introduced into the steam drum behind a baffle where, because

the density of the cold water is greater, it descends in the 'down comer' towards the

lower or 'mud' drum, displacing the warmer water up into the front tubes.

Continued heating creates steam bubbles in the front tubes, which are naturally sep-

arated from the hot water in the steam drum, and are taken off.

Water tube boilers support higher pressure (up to 35 MPa) and temperature (900 K) than

fire-tube boilers, but are more complex, larger (up to 50 m high, up to 60 kg/s of steam), and

more expensive than fire-tube boilers. In supercritical boilers, water is heated at more than

22 MPa and converted to supercritical steam without any phase change.

Figure 2-8 Water tube boiler schematic


25/81

13

Here’s a comparison table of some aspects of water tube and fire tube boilers:

Figure 2-9 Comparison between fire tube and water tube boilers

2.3.8 Waste Heat Recovery Boilers

Heat recovery boilers, also known as waste heat recovery boilers or heat recovery steam

generators (HRSGs), form an inevitable part of chemical plants, refineries, power plants, and

process systems. They are classified in several ways, as can be seen in Figure 2.10, according

to the application, the type of boiler used, whether the flue gas is used for process or mainly

for energy recovery, cleanliness of the gas, and boiler configuration, to mention a few. The

main classification is based on whether the boiler is used for process purposes or for energy

recovery. Process waste heat boilers are used to cool waste gas streams from a given inlet

temperature to a desired exit temperature for further processing purposes. Steam genera-

tion is of secondary importance in such plants. In energy recovery applications, on the other

hand, the gas is cooled as much as possible in order to obtain the highest possible steam

quantity.


26/81

14

Figure 2-10 classification of waste heat boilers

Getting back to our project, and taking into account our system considerations that will bediscussed later, we’ve decided to design a waste heat recovery fire tube boiler since it’s

easier to manufacture and costs less than water tube boiler. In addition, fire tubes are

used usually to produce steam at small scale.

The fire tube boiler has the following criteria based on Figure 2-10 above:

Gas condition: Dirty

Purpose: Energy Recovery from flue gases

Circulation: Natural

Firing: Unfired

Steam System: Single Pressure, Saturated Steam

Configuration: Single Pass, Integral Drum


27/81

15

3 Project Data

3.1 Project Description

As explained before, the chronic electricity shortage in Lebanon and particularly the fre-

quent power outages that occur on daily basis has led the industrial sector to generate its

own electrical power using Diesel Engine Generators which incur a heavy financial burden on

the factories especially after the rise in the price of fuel over the last decade and resulting to

the increase of their products’ market prices.

As a matter of fact, these frequent power outages have a very undesirable impact on the

factory’s production line and cause the production chain to stop for some time as well as the

production of non-completed products and putting many delicate and relatively expensive

systems in risk of failure. All this impose meaningful expenses and loss of productivity in the

factory.

For instance, the POPPINS® cornflakes factory had established a generator’s room next tothe factory in order to generate its own power needs.

Six CATERPILLAR® Diesel Engine Generators are connected in parallel in the generators

room and each generator set has the following specs:

Model: CAT® 3412C Diesel Engine

Power Rating @ 0.8 power factor : 725 kVA

50 Hz, 1500 rpm, 400 Volts

Figure 3-1 Caterpillar Diesel Engine


28/81

16

Figure 3-2 Diesel Engine Specifications

On the other hand, some food processes in the factory require a large amount of steam.

Particularly, the steam is used to cook the cornflakes. Thus, a steam boiler is installed in the

factory and is able to produce the required amount of steam used in the food process using

fuel oil as its main power source.

Pressure sensors throughout the steam network provide the sufficient feedback data to the

fuel oil burners in order to maintain the required steam pressure at 6 bars. However, the

increasing global fuel prices made the production of steam a bit more expensive.

Facing these two issues, the company decided to somehow merge the two systems and

solve the problem. In fact, the heat from the Diesel Generators exhaust gas will be used to

produce steam in a smaller unit.


29/81

17

The final system design should be able to extract the heat out of the exhaust gas using a

waste heat fire tube boiler in order to produce a fair amount of pressurized steam that will

be connected to the main steam network at the factory. ( see the scheme in Chapter 8 )

As a result:

The price of steam production will be reduced.

The overall efficiency of the Diesel Generators will be improved.

We note that the boiler should impose a low back pressure on the exhaust system in order

to prevent motor failure. This issue will be discussed in the next chapter.

The waste heat boiler will be put on top of the generators’ room and connected to the ex-

haust pipe from one side. Figures 3-3, 3-4 and 3-5 give us an idea about the boiler’s

placement:

Figure 3-3 Boiler placement


30/81


31/81

19

3.2 Input Data

In this section we will list the different system Input design data that are mainly the exhaust

gas physical and thermal properties.

As a matter of fact, a Diesel engine has fuel efficiency of about 40%, and around 30% of theenergy is transferred to the motor cooling fluid. That leaves around 30% of energy lost in the

atmosphere through the exhaust gas.

3.2.1 Exhaust Gas

First, if we examine the CAT® Genset Diesel Engine technical data sheet under “Exhaust Sys-

tem” (figure 3-6).

Figure 3-6 Diesel Engine Technical Data


32/81

20

We conclude the following valuable data:

Exhaust stack gas temperature: 534 ˚C (right on the motor exit)

400 ˚C (measured on the roof)

Exhaust gas flow rate: 125.4 m3/min

Total heat rejection to exhaust: 571 kW

Exhaust system max allowable backpressure: 6.7 kPa

Exhaust gas pressure : 1 atm

3.2.1.1 Temperature

The exhaust gas temperature is given right after the exhaust leaves the combustion chamber

into the exhaust system. Therefore the actual exhaust temperature to be used in our designis less than 534 ˚C due to the heat loss through the exhaust pipe walls, and in order to know

its correct value, we’ve managed to use a temperature measuring device and found out that

the actual temperature of flue gas on the roof is around 400 ˚C. This would be our gas design

temperature.

3.2.1.2 Composition

Generally, flue gases obtained from Diesel combustion contain a mixture of the following

components grouped in the table 3 below:

Major Constituents Minor Constituents (less than 1%)

Nitrogen, N2 Sulfur Oxides, SO2, SO3Water Vapor, H2O Nitrogen Oxides, NO, NO2

Carbon Dioxide, CO2 Carbon Monoxide, CO

Oxygen, O2 Hydrogen, H2

Table 3 Exhaust gas major and minor constituents

In fact, the presence of gases such as hydrogen and water vapor increases the heat transfer

coefficient significantly, which can affect the heat transfer and the boiler size. Also, if the gas

is at high pressure, the mass velocity inside the tubes can be much higher because of the

higher density, which also contributes to the higher heat transfer coefficients. However, due

to the absence of exact gas composition data we will assume that:

The latent enthalpy recovery due to the condensation of vapor present in the flue gas

will not be taken into consideration in the design.

The exhaust gas composition:

N2 = 76%

CO2= 13%

Water vapor H2O=11%


33/81


34/81


35/81


36/81

24

Table 6 - Saturated steam pressure table

3.4 Preliminary Calculation

The preliminary calculations were made as a first step to get an idea on whether the process

of generating fair amounts of steam is possible and know the amplitude of the power that

we’re dealing with.

For instance, let’s assume that the gas will be cooled from:

To

Then


37/81

25

Where: is the average temperature of the gas between the inlet and the outlet.First, knowing the gas flow volumetric rate, the total mass flow rate of the gas must be de-

termined from the equation:

̇ ̇ (1)Where ̇ ̇

For

, we interpolate the value of the gas density from table 4:

And the specific heat capacity of the gas:

Then from equation (1),

̇

̇ Next, we calculate the heat load available from the flow of gas:

̇

Finally,

On the other hand, if we assume that we’ve managed to transfer the entire heat load from

the gas to the water, without any heat exchanger calculations, then we can calculate the

flow of steam that we can produce and this could give us an idea about the quantities we’re

dealing with and whether the system is practical or not.

On the waters side:

̇ (2)


38/81

26

Where

̇

Then for we obtain:̇ ̇ Interestingly, we conclude that we can produce a fair amount of steam out of the waste

heat. Thus, the project is valuable to the factory, and we’ll proceed to the next step which is

to design the fire tube waste heat recovery boiler.

The design should take into consideration the following:

o Heat transfer design and sizing.

o Off-design performance.

o Pressure drop calculations.

o Sizing of the pressure vessel and the tube sheets.

o Manufacturing considerations.

o Boiler accessories (safety valves, water level meter, feedwater pump, etc...).


39/81


40/81

28

Thus, Eq. 3 a can be rewritten as either

(3.a)

Or

(3.b)The energy transferred per unit time, Q, is:

[ ̇ ] ̇ (5)The term r represents the heat loss factor and is equal to one minus the losses due to radia-

tion and convection from the boiler surfaces. A 2% loss, or r = 0.98 , is typical

The log mean temperature difference, is determined by:

(6)

The overall heat transfer coefficient Uo is given by:

(7)Where

. .

Where R f is the fouling resistance.

The value of hi is obtained from the Dittus-Boelter correlation:

(8)For 0.6


41/81

29

(10) V is the velocity of the gas inside the tubes and is the kinematic viscosity of flue gas.The 3rd term in Eq. 7 is the resistance of the tube wall to heat transfer. The thermal conduc-

tivity of the tube material, K, is about 35-45 W/m-K for carbon steel, the typical material

used for boilers.

To size the boiler, the mass flow per tube, ranging from 50 to 90 kg/hr. for a 2-in. tube, and

the gas velocity, typically ranging from 20 to 50 m/s, are assumed and the tube count is cal-

culated. The relationship between mass flow and velocity is:

̇ (11)In practice, it’s easier to assume a number of tubes and choose the tubes ID first then calc u-

late the gas velocity inside it.

Based on the temperature and properties of the gas, all the variables are calculated and fi-

nally U is determined.

Then Eq. 3 is used to calculate A which in turn used to determine the tube length, L.

At the end, the pressure drop is calculated (see section 4.3), and if the computed pressure

drop is higher than that allowed by the specification, another tube count or mass flow rate

per tube is assumed and the procedure is repeated. [11]

The flowchart in figure 4-1 will visually explain the steps that have been made to size the fire

tube boiler.

Now regarding the first step of our sizing procedure, we’ve assumed the following:

We must size the boiler to be able to cool the exhaust gases from 400˚C to 230 ˚C.

OD 1.75”; 1.521”ID steel tubes were chosen.(available in the market for 6 meters of

length )

An average of 3.5 generators running all the time.

A tube count of 200 will be assumed as a start and will be checked in the end for

pressure drop and will be optimized for better heat transfer, n=200.


42/81

30

Choose d i

Assume tube count n

Calculate ̇ per tube45


43/81

31

4.2 Numerical Applications

4.2.1 Gas and steam properties

The boiler’s size should be able to cool the exhaust gas from:

To Thus, all the gas properties will be determined from Table 4 for the gas mean temperature:

All the needed exhaust gas properties are listed below:

⁄ As for the feedwater and steam, we’ll only use the enthalpies at feedwater and steam at

saturation temperatures obtained from Tables 5 and 6:

4.2.2 Mass flow per tube

First, knowing the gas flow volumetric rate, the total mass flow rate of the gas must be de-

termined from the equation:

̇ ̇

Where

̇ ̇ Then

̇


44/81

32

̇ And the mass flow per tube is acceptable, it is given by:

̇ ̇ ̇ 4.2.3 Exhaust Gas Velocity

As we will discuss later, the exhaust gas velocity inside the tubes has an important role in

increasing the heat transfer between the gas and tubes.

We note that the tubes chosen are 1.75”-1.521” tubes, that’s the equivalent of:

And From eq. 11:

4.2.4 The Reynolds NumberIn fluid mechanics, the Reynolds number (Re) is a dimensionless number that gives a meas-

ure of the ratio of inertial forces (which characterize how much a particular fluid resists any

change in motion) to viscous forces and consequently quantifies the relative importance of

these two types of forces for given flow conditions.

It is often used to characterize different flow regimes, such as laminar or turbulent flow:

laminar flow occurs at low Reynolds numbers, where viscous forces are dominant, and ischaracterized by smooth, constant fluid motion; turbulent flow occurs at high Reynolds

numbers and is dominated by inertial forces, which tend to produce chaotic eddies, vortices

and other flow instabilities.

In our case, the Reynolds number is used to calculate the heat transfer coefficient between

the gas side and the tubes. The higher the Reynolds number, the higher the heat transfer.

The Reynolds number is calculated below:


45/81

33

4.2.5 The Nusselt Number and Dittus-Boelter Correlation

The Dittus-Boelter equation (for turbulent flow) is an explicit function for calculating the

Nusselt number. It is easy to solve but is less accurate when there is a large temperature

difference across the fluid. It is tailored to smooth tubes, so use for rough tubes (most com-

mercial applications) is cautioned. The Dittus-Boelter equation (Eq. 8) can be used sinceRe>10000, Pr>0.6 and L/D>10:

4.2.6 Gas side heat transfer coefficient hi

The gas side heat transfer coefficient hi can be derived from Eq. 9 now that we’ve calculated

the Nusselt number:

⁄ ⁄

We will conclude later that hi dominates the Overall heat exchange coefficient U . In fact, the

sum of all the other resistances will be neglected because they only contribute in 8% only of

the U value.

4.2.7 Water side heat transfer coefficient ho

The boiling heat-transfer coefficient ho is very high - on the order of 3000 to 100 000

W/m2.K. Thus, even a 20% variation in its value will not impact U, because the tube-side co-

efficient, hi , which is typically on the order of 50-100 W/m2.K, governs U .

However, before the nucleate boiling starts we should make sure that ho is still relatively

high and does not affect the U value that much. We should calculate the natural convection

heat transfer coefficient.

In fact, convection heat transfer takes place when a fluid flows past a solid surface, with a

difference in temperature between the fluid and the surface. If the fluid flow is due to an

external force, like a pump or fan, it is forced convection. If the fluid flow is caused by densi-

ty differences within the fluid due to internal fluid temperature differences, then it is natural

convection, also sometimes called free convection.

The equations used to calculate natural convection heat transfer coefficients come from

correlations of dimensionless numbers. The dimensionless numbers typically appearing in

these correlations are the Nusselt number, the Prandtl number, the Grashof number, and

sometimes the Rayleigh number. The equations for the Nusselt, Prandtl, and Grashof num-bers (Nu, Pr, and Gr) are shown in the box below. The Rayleigh number is simply: Ra = Gr Pr.

http://img.bhs4.com/63/d/63da4aaa40615498f8a138e7c720f9c6dd5c92ed_large.jpghttp://img.bhs4.com/63/d/63da4aaa40615498f8a138e7c720f9c6dd5c92ed_large.jpg


46/81

34

Figure 4-2 Nusselt, Grashof and Prandtl numbers

Following are the parameters that appear in these dimensionless numbers:

D is a characteristic length parameter (e.g. diameter for natural convection from a cir-

cular cylinder or a sphere or height of a vertical plate) in m.

ρ is the density of the fluid in Kg/m3 .

μ is the kinematic viscosity of the fluid N-s/m2 .

k is the thermal conductivity of the fluid W/m-K.

C p is the heat capacity of the fluid in J/kg-K.

g is the acceleration due to gravity (9.81 m/s2).

β is the coefficient of volume expansion of the fluid in K -1.

ΔT is the temperature difference between the solid surface and the fluid.

In our case, it’s a natural convection heat transfer from a horizontal tube outside surface and

the water.

The Nusselt number/Rayleigh number/Prandtl number correlation for natural convection heattransfer between a fluid and an isothermal horizontal cylinder is shown in figure below. An

Excel spreadsheet was made in order to easily calculate ho .

Figure 4-3 Natural convection heat transfer from an isothermal horizontal cylinder


47/81

35

As shown in the equations in the box, figure 4-2, the length parameter used in the Nusselt

number and Grashof number is the cylinder diameter, D. There is a single correlation for the

Nusselt number for this configuration. It applies for Rayleigh number less than 1012.

We note that the water properties were extracted from the pressurized water tables at 6 bars

pressure.

The figure 4-3 above shows the calculations made in the excel spreadsheet available on the

soft copy. It clearly shows how large is ho , thus it can be neglected in the calculations of U.

Figure 4-4 Excel spreadsheet calculations for natural convection


48/81


49/81

37

4.2.9 Log mean temperature difference ΔTLM

The log mean temperature difference (also known by LMTD) is used to determine the tem-

perature driving force for heat transfer in the fire tube boiler since the entrance and exit

temperatures of the fluids are not the same.

In our case:

4.2.10 Boiler Duty Q

The boiler duty Q represents the heat load available from cooling the gas from 400 ˚C to230˚C; Eq. 5:

̇ Hence

4.2.11

Surface Area and Tube Length

The heat exchanger’s surface area is obtained from the equation 3.b :

And finally we deduce the length of the tubes:

Now that we’ve calculated the dimension of the boiler, we must verify the design for pres-

sure drop.


50/81

38

4.3 Pressure Drop

Back pressure refers to pressure opposed to the desired flow of a fluid in a confined place

such as a pipe. It is often caused by obstructions or tight bends such as piping or air vents.

Because it is really resistance, the term back pressure is misleading as the pressure remainsand causes flow in the same direction, but the flow is reduced due to resistance.

Back pressure caused by the exhaust system of an engine has a negative effect on engine

efficiency resulting in a decrease of power output that must be compensated by increasing

fuel consumption, and finally, if the backpressure is high, it will cause the diesel engine to

shut down.

Our goal now is to determine the pressure drop of the exhaust gas across the fire tube boiler

and make sure that it is not beyond its limit.

Mainly, pressure drop is the result of frictional forces on the fluid as it flows in the core of

the tube.

But first, a fluid experiences an entrance loss as it enters the tubes due to a sudden reduc-

tion in flow area as well as it exits the core due to a sudden expansion.

In fluid dynamics, the Darcy –Weisbach equation relates the pressure loss due to friction

along a given length of pipe to the average velocity of the fluid flow.

The Darcy –Weisbach equation contains a dimensionless friction factor, known as the Darcy

friction factor. This is also called the Darcy –Weisbach friction factor or Moody friction factor.

The expression for turbulent flow pressure drop of fluids (Reynolds number >2100) is:

(12) Where


51/81


52/81

40

Before choosing a formula to calculate f , it is worth knowing that in the paper on the Moody

chart, Moody stated the accuracy is about ±5% for smooth pipes and ±10% for rough pipes.

We note that the absolute roughness of the tubes made of commercial steel is:

So the relative roughness is:

From the Moody chart we conclude that:

However, although the Moody chart is accurate, it is also impractical to use and an analytical

method is preferred.

Colebrook equation

The Colebrook equation is an implicit equation that combines experimental results of studies

of turbulent flow in smooth and rough pipes. The equation is used to iteratively solve for the

Darcy –Weisbach friction factor f . This equation is also known as the Colebrook –White equa-

tion.

However, this equation is impractical and needs a couple of iterations in order to converge

to the exact value of f.

Swamee –Jain equation

The Swamee –Jain equation is used to solve directly for the Darcy –Weisbach friction factor f

for a full-flowing circular pipe. It is an approximation of the implicit Colebrook –White equa-tion.

The Swamee-Jain equation is explicit and easy to solve, and it gives a relatively accurate and

tolerable result.

The calculation is implemented in the spreadsheet and described below:


53/81


54/81

42

4.3.2.1 Entrance Loss

Figure 4-5 Entrance flow conditions and loss coefficient (a) Reentrant, K=0.8, (b) sharp-edged, K=0.5,

(c) slightly rounded, K=0.2, (d) well-rounded, K=0.04.

4.3.2.2 Exit Loss

Figure 4-6 Exit flow conditions and loss coefficient (a) Reentrant, K=1, (b) sharp-edged, K=1, (c) slight-

ly rounded, K=1, (d) well-rounded, K=1.


55/81

43

4.3.2.3 Equivalent Length

Considering the tube to tube sheet welding technique, we can tolerably assume that the

entrance and exit of the tubes are both sharp-edged, then:

And finally,

( )

4.3.3 Pressure Drop

The pressure drop of the gas across the fire tube boiler is finally given by Eq. 12:

We notice that:

Hence, the [tube length/tube count/tube ID] configuration that we’ve considered is valid

and causes allowable backpressure to the Diesel engines.

Furthermore, similar calculations were made (Excel Spreadsheet) considering that the 6 en-gines are all running at full power at the same time instead of 3.5 which is in fact the most

critical case for backpressure.

For 6 engines running

Yet, several other configurations can produce a better heat transfer and still an allowable

pressure drop. This point will be discussed in the next section.


56/81

44

4.3.4 Heat Transfer vs. Pressure Drop

Obviously, different boiler configuration may lead to different heat transfer and pressure

drop.

First, let us examine the following table where the values were calculated using the excel

spreadsheet:Size, in 1.75x1.521 2.0x1.77 2.5x2.238

Number of tubes 150 200 250 130 180 230 100 150 200

Velocity, m/s 41.6 31.2 24.96 35.44 25.6 20.03 28.82 19.21 14.41

Length, m 3.67 3.46 3.31 4.26 4.0 3.8 5.42 5 4.71

Surface Area, m2 76.83 96.72 115.62 88.42 114.71 139.56 108.1 149.51 188.2

Uo, W/m2-K 77.37 61.47 51.42 67.24 51.83 42.6 55.0 39.76 31.58

ΔP, kPa 2.12 1.15 0.72 1.54 0.77 0.46 1.02 0.43 0.23

Table 7 - Results of design calculations for fire tube waste heat boilers for the same duty

Surface area should not be used as the sole criterion for selecting boilers, because tube size

and gas velocity affect this variable.

Shown in table 7 are the design options for the same boiler duty using different gas veloci-

ties and tube sizes; the procedure described in the last section was used to arrive at these

options. The purpose behind this table is to bring out the fact that surface area can vary byas much as 50% for the same duty.

1. As the gas velocity increases, the surface area required decreases, which is obvious.

2. The smaller the tubes, the higher the heat transfer coefficient for the same gas ve-

locity, which also decreases the surface area.

3. For the same gas pressure drop, the tube length is smaller if the tube size is smaller.

This fact helps when we try to fit a boiler into a small space.

4. For the same tube size, increasing the gas velocity results in a longer boiler, a greater

gas pressure drop, but smaller surface area.

So is surface area an important criterion for evaluating different boiler designs?

The answer is yes if the person evaluating the designs is knowledgeable in heat transfer –

related aspects and no if the person simply compares different designs looking only for sur-

face area information. We can observe this in the Table 7 which shows the results of design

calculations for fire tube waste heat boilers in different configurations, and where, due to

variations in tube size and gas velocity, different designs with over 40 –50% difference in sur-

face areas were obtained for the same duty Q.

The interpretation of these values can be concluded in the following diagrams:


57/81

45

Need to increase heat transfer?

As a matter of fact, when the surface area increases, the heat transfer increases.

Similarly, increasing the gas velocity inside the tube increases the heat transfer coefficient hi

and therefore the heat transfer increases.

Figure 4-7 Influence of various geometrical parameters of a shell-and-tube exchanger on heat trans-

fer and pressure drop.

Need to increaseheat transfer

Increase heattransfer

coefficient

Tube Side

Increase numberof tubes

Decrease tubeoutside Diameter

Shell Sidenot applicable infire tube boilers

Increase surfacearea

Increase tubelength

Increase shelldiameter and

number of tubes

Employ multipleshells in series or

parallel


58/81

46

Need to reduce pressure drop?

The pressure drop is a function of the gas velocity, V , and the length of the tube, L.

Increasing the length will increase the pressure drop.

Increasing the tube diameter or the number of tubes will decrease V and we obtain a lower

pressure drop.

Figure 4-8 Influence of various geometrical parameters of a shell-and-tube exchanger on heat trans-

fer and pressure drop.

Need to reduce

pressure drop

Tube Side

Decrease number oftube passes

Increase tube diameter

Decrease tube lengthand increase shell

diameter and numberof tubes

Shell SideNot applicable in fire

tube boilers


59/81


60/81

48

Known:

Tube length

Tube diameter

Number of tubes

Assume:

Gas exit temperature

t2

Determine Gas properties

Calculate Gas Velocity

Calculate Reynolds number

Calculate Nusselt number

Calculate U

Deduce hi

̇ Calculate Duty Q

̇ Calculate steam genera-

tion:

̇ Solve for T2 unknown: ?N

END

Y

Figure 5-1 Flow chart for fire tube design procedure


61/81

49

The boiler duty Q is given by the expression:

(13)

̇ Simplifying, we have:

̇ (14) We assume that t 2= 230˚C as the gas exit temperature, then

At this temperature, the gas properties are the following:

⁄ The internal velocity: Gas mass flow rate: ̇ Reynolds number:

Nusselt number: Convection coefficient : ⁄ U-Value:

⁄


62/81


63/81

51

Hence,

This time we get:

Conclusively, the two values have converged and we can now predict that the exit tempera-

ture of the exhaust gas is:

Hence, we can now calculate the duty of the boiler and the flow rate of the steam generat-

ed:

̇ ̇

̇ The calculations procedure is implemented in the excel spreadsheet, but manual iteration of

the values is needed.

Heat Exchanger Effectiveness Є

Effectiveness Є is a measure of thermal performance of a heat exchanger. It is defined for a

given heat exchanger of any flow arrangement as a ratio of the actual heat transfer rate

from the hot fluid to the cold fluid to the maximum possible heat transfer rate Q max thermo-

dynamically permitted:

Or Hence,


64/81

52

6 Boiler’s Mechanical Design

6.1 Pressure vessel

6.1.1 Introduction

The pressure vessels (i.e. cylinder or tanks) are used to store fluids under pressure. The fluid

being stored may undergo a change of state inside the pressure vessel as in case of steam

boilers or it may combine with other reagents as in a chemical plant. The pressure vessels

are designed with great care because rupture of pressure vessels means an explosion which

may cause loss of life and property. The material of pressure vessels may be brittle such that

cast iron or ductile such as mild steel.

Cylindrical or spherical pressure vessels (e.g., hydraulic cylinders, gun barrels, pipes, boilers

and tanks) are commonly used in industry to carry both liquids and gases under pressure.

When the pressure vessel is exposed to this pressure, the material comprising the vessel is

subjected to pressure loading, and hence stresses, from all directions. The normal stresses

resulting from this pressure are functions of the radius of the element under consideration,

the shape of the pressure vessel (i.e., open ended cylinder, closed end cylinder, or sphere) as

well as the applied pressure.

Two types of analysis are commonly applied to pressure vessels. The most common method

is based on a simple mechanics approach and is applicable to “thin wall” pressure vessels

which by definition have a ratio of inner radius, r, to wall thickness, t, of r/t≥10. The secondmethod is based on elasticity solution and is always applicable regardless of the r/t ratio and

can be referred to as the solution for “thick wall” pressure vessels.

In our analysis we will discuss the thin wall pressure vessel approach in order to determine

the type and thickness of the material that forms the body of our cylindrical fire tube boiler

according to the ASME code (American Association of Mechanical Engineers.)

6.1.2 Manufacturing Constraints

While perforating the tube sheets, a minimum inter-tube distance of 3 cm should be givenaccording to the manufacturer. This will affect the diameter of the shell and the tube sheet.

6.1.3 Shell Dimensions

Given:

An AutoCad drawing (Figure 6-1) was used to determine the tube sheet diameter. The boiler

must have the following dimensions:


65/81

53

Figure 6-1 Tubesheet AutoCad drawing and dimensions

6.1.4 Material and Maximum Allowable Stress

Mild steel such as SAE-285 Grade C steel which is a typical hot rolled mild steel [4] will be

used in the construction of the pressure vessel.

According to the ASME code, the design should be made with respect to the Maximum Al-

lowable Stress of the material and not its Yield strength unlike European codes.

The maximum allowable stress value is the maximum unit stress permitted in a given mate-

rial used in a vessel constructed under the ASME code [5] It is usually based on 2/3rd of theYield strength of the material (equivalent to a 1.5 safety factor in Euronorm). In the design of

a pressure vessel the max allowable stress is based on one-fourth of the ultimate tensile

strength of the material.(equivalent to a 4 safety factor in Euronorm) [4].

The hazard of an exploding vessel is great, a fact which justifies the use of a greater factor of

safety for pressure vessels than any other application.

The max allowable stress for SAE-285 Grade C steel temperature not exceeding 400 ˚F [6]

(204˚C) is:


66/81

54

6.1.5 Loadings and Design Pressure

Vessels covered by this Division of Section VIII shall be designed for at least the most severe

condition of coincident pressure and temperature expected in normal operation. For this

condition the maximum difference in pressure between the inside and outside of a vessel.

[7]

The loadings to be considered in designing a vessel shall include those from:

6.1.5.1 Internal and external pressure

6.1.5.2 Weight of the vessel and normal contents under operating or test conditions

This includes additional pressure due to static head of liquids that is maximum at the bottom

of the vessel:

Where

Finally

6.1.5.3 The final design pressure

The final inside design pressure shall be:


67/81

55

6.1.6 Pressure Vessel thickness under internal pressure

The thickness of shells under internal pressure shall be not less than that computed by the

following formulas. The symbols defined below are used in the formulas of this paragraph:

. P P

For cylindrical shells, the minimum thickness should be greater than the greatest thickness

given by the formulas below [8]:

1. Circumferential Stress (Longitudinal joints)

2. Longitudinal Stress (Circumferential joints)

Figure 6-2 explains the type of stresses in a cylindrical shell:

Figure 6-2 types of stresses in a cylindrical shell, S1=longitudinal stress, S2=Circumferential or hoop

stress


68/81

56

Numerical application:

1. For circumferential stress,

2. For longitudinal joints, Finally,

And for a corrosion allowance of C.A. = 1mm

6.2 Tube Sheet Design

The tube-plates (tube-sheets) in shell and tube heat exchangers support the tubes, and sep-

arate the shell and tube side fluids. One side is subject to the shell side pressure and the

other the tube-side pressure. The plates must be designed to support the maximum differ-

ential pressure that is likely to occur. Radial and tangential bending stresses will be induced

in the plate by the pressure load and, for fixed-head exchangers, by the load due to the dif-

ferential expansion of the shell and tubes.

A tube-plate is essentially a perforated plate with an imperforated rim, supported at its pe-

riphery. The tube holes weaken the plate and reduce its flexural rigidity.

This chapter discusses the design of fixed tube sheets in accordance with the method pro-

posed by Dr. K. A. G. Miller. It takes into account the support given to the tube sheets by thetubes and also the weakening effects of different tube hole spacing. The tube sheet designed

by this method results in thickness much less than as given by the method proposed by

TEMA (Tubular Exchanger Manufacturers Association).

The Miller method is generally preferred over the TEMA method for economical purposes,

especially for large diameter alloy tube sheets designed for low internal pressure. There, will

not only be a saving material but, more important, a saving in the machining time for drilling

the holes in the tube sheet.


69/81

57

6.2.1 Design Procedure

Figure 6-3 Fire tube boiler scheme

Cross-sectional area of one tube is:

Cross-sectional area of inside of shell is:

Cross-sectional area of tube holes in tube sheet is given by:

Cross-sectional area of shell plate is found using the formula:

Ligament efficiency can be calculated from the relationship:

Determine:


70/81

58

6.2.2 Working Conditions

Calculate equivalent pressure difference by:

Differential thermal expansion is:

Effective pressure difference due to the combined pressure difference P and the differential

expansion is:

Determine the value of dimensionless factor:

kR G1 G2 G3 G4

0 0.800 0.800 +1.000 1.000

0.5 0.809 0.810 +0.998 1.002

1.0 0.820 0.844 +0.966 1.0291.5 0.871 0.993 +0.836 1.14

2.0 1.012 1.412 +0.546 1.40

2.5 1.34 2.40 +0.121 1.79

3.0 1.88 4.24 -0.306 2.25

3.5 2.36 6.36 -0.608 2.69

4.0 2.75 8.53 -0.741 3.10

4.5 3.10 10.75 -0.727 3.47

2.0 3.43 13.1 -0.619 3.83

5.5 3.77 15.8 -0.541 4.18

6.0 4.12 18.7 -0.515 4.54

7.0 4.82 25.3 -0.529 5.26

8.0 5.54 33.1 -0.564 5.97

9.0 6.26 41.8 -0.602 6.68

10.0 6.98 51.6 -0.642 7.39

12.0 8.43 74.3 -0.727 8.81

14.0 9.88 101.1 -0.816 10.23

16.0 11.33 132.0 -0.907 11.65

18.0 12.80 167.2 -0.999 13.06

20.0 14.25 206.4 -1.091 14.48

Table 8 Values for G1, G2, G3 and G4


71/81

59

The values of G1, G2, G3 and G4 corresponding to the factor kR can be read from Table 8.

Maximum radial stress in tube plate is given by:

* + () Also, maximum stress in tube material is greater of:

( )

Or

( )

If, either of the stresses in any of the cases is found more than the allowable, the tube plate

thickness should be modified unless the stresses within allowable limits are obtained. [9]

6.2.3

Numerical Application

We note that ASTM A53 type F Grade B steel is one of the widely used pipe and tube materi-

al and there is no need to check the tube thickness for failure because small diameter tubes

can withstand very high internal and external pressures.

We also note that the tube sheet is made of the same material as the vessel, SAE-285 Grade

C steel.

Now we gathered the following data:

Design temperature around 200˚C or 400˚F

, the shell design temperature. , Shell bore.

, Modulus of elasticity of shell material [10]


72/81

60

⁄ , Coefficient of thermal expansion for shell

, the tube design temperature. , Modulus of elasticity of tube material

⁄ , Coefficient of thermal expansion for tubes

, Modulus of elasticity of tube sheet materialAssuming the total thickness of tubesheets as 2 in. or 50.8 mm therefore,


73/81

61

The effective length is:

From Table 8, for kR=4.13 we get by interpolation:

Hence,

* +

(

)

And

( )

Or

( )

Since all the stresses are within allowable limits, a 2 inch thick tubesheet is sufficient for this

exchanger. Thickness could be further reduced but seems to be quite reasonable and safefor such an exchanger.


74/81

62

7 Boiler fittings and mountings

A number of items must be fitted to steam boilers, all with the objective of improving [12]:

Operation.

Efficiency.

Safety.

7.1 Safety valves

An important boiler fitting is the safety valve. Its function is to protect the boiler shell from

over pressure and subsequent explosion.

In Europe, matters relating to the suitability of safety valves for steam boilers are governedby the European standard EN 12953. In the US and some other parts of the world, such mat-

ters are covered by ASME standards.

Many different types of safety valves are fitted to steam boiler plant, but generally they

must all meet the following criteria:

The total discharge capacity of the safety valve(s) must be at least equal to the 'from

and at 100°C' capacity of the boiler. If the 'from and at' evaporation is used to size

the safety valve, the safety valve capacity will always be higher than the actual max-

imum evaporative boiler capacity.

The full rated discharge capacity of the safety valve(s) must be achieved within 110%

of the boiler design pressure.

The minimum inlet bore of a safety valve connected to a boiler shall be 20 mm.

The maximum set pressure of the safety valve shall be the design (or maximum per-

missible working pressure) of the boiler.

There must be an adequate margin between the normal operating pressure of the

boiler and the set pressure of the safety valve.

Figure 7-1 Boiler safety valve


75/81

63

7.2 Boiler stop valves

Figure 7-2 Boiler stop valve

A steam boiler must be fitted with a

stop valve (also known as a crown

valve) which isolates the steam boiler

and its pressure from the steam net-

work. It is generally an angle pattern

globe valve of the screw-down varie-

ty. Figure 8-2 shows a typical stop

valve of this type.

In the past, these valves have often been manufactured from cast iron, with steel and

bronze being used for higher pressure applications. In the UK, BS 2790 (eventually to be re-

placed with EN 12953) states that cast iron valves are no longer permitted for this

application on steam boilers. Nodular or spheroidal graphite (SG) iron should not be con-

fused with grey cast iron as it has mechanical properties approaching those of steel. For this

reason many boilermakers use SG iron valves as standard.

7.3 Feedwater Check Valve

The feedwater check valve (as shown in Figures 8-3 and 8-4) is installed in the boiler feedwa-

ter line between the feedpump and boiler. A boiler feed stop valve is fitted at the boiler

shell.

The check valve includes a spring equivalent to the head of water in the elevated boiler

when there is no pressure in the boiler. This prevents the feed pump being flooded by the

static head from the boiler.

Figure 7-3 Boiler Check Valve

Under normal steaming conditions the

check valve operates in a conventional

manner to stop return flow from the boiler

entering the feed line when the feedpump

is not running. When the feedpump is run-

ning, its pressure overcomes the spring to

feed the boiler as normal.


76/81

64

Figure 7-4 Location of feed check valve

7.4 Pressure gauge

All boilers must be fitted with at least one pressure indicator. The usual type is a simple

pressure gauge constructed to EN 12953.

The dial should be at least 150 mm in diameter and of the Bourdon tube type, it should be

marked to indicate the normal working pressure and the maximum permissible working

pressure / design pressure.

Pressure gauges are connected to the steam space of the boiler and usually have a ring typesiphon tube which fills with condensed steam and protects the dial mechanism from high

temperatures.

Figure 7-5 Typical pressure gauge with ring siphon


77/81

65

7.5 Water Level Indicator

All steam boilers are fitted with at least one water level indicator, but those with a rating of

100 kW or more should be fitted with two indicators. The indicators are usually referred to

as gauge glasses complying with EN 12953.

Figure 7-6 Gauge glass and fittings

A gauge glass shows the current level of water in the boiler, regardless of the boiler's operat-

ing conditions. Figure 8-6 shows a typical gauge glass.

7.6 Water level controls

The maintenance of the correct water level in a steam boiler is essential to its safe and effi-

cient operation. The methods of sensing the water level, and the subsequent control of

water level is a complex topic that is covered by a number of regulations. The t

Documents

Ghinwa JASSAR-Jonny Saidy.pdf