Mechanical Design [Compile Draft 1]

1

CHAPTER III

MECHANICAL DESIGN OF EQUIPMENT

3.1 REACTOR 1, R1

3.1.1 INTRODUCTION

In the mechanical design of process equipment, there are many aspects of

design and reactor safety factors should be considered. Among these is the stress

analysis, the burdens imposed on the reactor and the reactor design supporters. All

these aspects are based on a standard code of the American Society of Mechanical

Engineers (ASME).

Tube and shell reactor was operated in the gas phase and liquid phase at a

temperature of 185 0C and pressure of 6.5 bar (650 kPa) design pressure, P took a

safety factor of 10% above the operating pressure.

3.1.2 MATERIAL OF CONSTRUCTION

Materials selection was based on the consideration of four main factors:

resistance to ammonia, nitric acid and ammonium nitrate vapours and condensate,

strength, ease of fabrication, and low cost. Much of the vessel (both the shell and

the tubes) will be in continuous contact with ammonium nitrate aqueous at high

temperatures. Therefore, particular attention was given to corrosion resistance

under those conditions. The tubes are in direct contact with both the cooling medium

and the reaction gases.

The preferred construction material for the reactor is stainless steel 16Cr-

2Mo-8Ni (316), which is described by the materials specification given in Table 5.2

and composition of material in Table 5.3. Ammonium nitrate, ammonia and nitric

acid are not particularly corrosive to most steels. The average corrosion rates are

generally less than 0.001 per year. The addition of chromium also improves the

mechanical properties at high temperature. Several stainless steels, notably type

316, satisfy all the material requirements. However, A387 is substantially cheaper

and can be used with little penalty to the corrosion rate. At high pressures (and,

2

hence, large wall thicknesses), cladding is normally recommended in order to

reduce the vessel cost when alloy steels are used.

Reactor construction material used is stainless steel 16Cr-2Mo-8Ni (316). By

referring to the standard code The American Society of Mechanical Engineers

(ASME), the maximum stress is 133.5 N/mm2.

By linear interpolation

Tem

perature, C 150 155 200

S,

N/mm2 135 S 120

Table 3.1 Typical mechanical properties for 316 stainless steel alloys

Property Value

Melting Point 1375-1400°C

Modulus of Elasticity 193 GPa

Electrical Resistivity 0.074x10-6 Ω.m

Thermal Conductivity 16.3 W/m.K at 100°C

Thermal Expansion 15.9x10-6 /K at 100°C

Tensile Strength (MPa) 515

Compression Strength (MPa) 170

Melting Point 1375-1400°C

Table 3.2 Typical chemical composition for 316 stainless steel alloys

% 316

C 0.08max

Mn 2.0

Si 0.75

P 0.045

S 0.03

Cr 16-18

3

Mo 2-3

Ni 10-14

N 0.1

3.1.3 THE EFFICIENCY OF WELDED JOINT

There are several methods to make welded joints. In particular case the

choices of a type from the numerous alternatives depend on:

1. The circumstances of welding.

In many cases the accessibility of the joint determines the types of welding.

In a small diameter vessel (under 18-24 inches) from the inside, no manual welding

can be applied. Using backing strip it must remain in place. In larger diameter

vessels if a manway is not used, the last (closing) joint can be welded from outside

only. The type of welding may be determined also by the equipment of the

manufacturer.

2. The requirements of the code.

Regarding the type of joint the Code establishes requirements based on

service, material and location of the welding. The welding processes that may be

used in the construction of vessels are also restricted by the Code as described in

paragraphUW-27.

3. The aspect economy.

If the two preceding factors allow free choice, then the aspect of economy

must be the deciding factor.

Some considerations concerning the economy of welding’s:

1. V-edge preparation, which can be made by torch cutting, is always more

econornical than the use of J or U preparation.

2. Double V preparation requires only half the deposited weld metal required for

single V preparation.

3. Increasing the size of fillet weld, its strength increases in direct proportion,

while the deposited weld metal increases with the square of its size.

4. Lower quality welding makes necessary the use of thicker plate for the

vessel. Whether using stronger welding and thinner plate or the opposite is

more economical, depends on the size of vessel, welding equipment, etc.

This must be decided in each particular case

The strength of a welded joint depends on the type and quality of welding

joint. Then, for design purposes weld joint efficiency, J = 1.0 was chosen. This

selection is based on ASME UW-2 stated that:

4

“… all butt welded joints shall be fully radiographe, except under provision

OS UW-2(a)(2) and UW-2(a)(3) below and UW-4(a)(4)….”

This statement is clarifying the requirement of welded joint that fully

radiograph when pressure vessel containing lethal substances. So, all main

category A and B welds must be fully radiographed. But category B and C welds in a

nozzle and communicating chambers that are not larger than 10 inch nominal pipe

size and do not exceed 1to 1/8 inch thick are exempt. Based on the fluid

composition contain in the reactor for this design, ammonium nitrate could be a

dangerous and lethal substance if leaking to the atmosphere. Furthermore,

ammonia also potentially dangerous substance. The location of A, B and C shown in

Figure 5.3.

Figure 5.3 Welded joint locations

3.1.4 DETERMINATION THICKNESS OF REACTOR SHELL AND HEAD

3.1.4.1 Design Pressure

From table 13.2 (R.K. Sinnot, 1999. Chemical Engineering Design), typical

design stress = 133.5 N/mm2

Operating pressure for reactor is 6.5 bar.

The pressure given in the table only design stress for selected material but

for design stress pressure that generated by the fluid also need to take into

consideration. From book of Pressure Vessel Handbook 10th edition page 29 giving

5

the pressure of water that will emit at different length. But for other material, the

value needs to multiply with specific gravity of fluid or other calculation is:

Value above is for the water. To get the pressure in the reactor emit by the

fluid is multiply value get by specific gravity of fluid. Specific gravity for the fluid in

the reactor is 0.1067.

So design pressure should be taken is:

Taking 10 per cent above as design pressure

3.1.4.2 Design temperature

Operating temperature = 185 0C

Take 10 percent above operating temperature,

3.1.4.3 Thickness of cylindrical vessel

Data required to performed calculation

Cylinder length, L = 15 m

Design pressure, P = 0.778 N/mm2

Inside diameter, Di = 4.80 m

Inside radius of reactor, R = 2.4 m

Allowable stress, S = 133.5 N/mm2

Joint efficiency, J = 1.0

Determination of reactor thickness, assume 0OD

t

for cylinder wall

i) Tangential stress with condition t < R/2 and P< 0.385SE

where

P = Design pressure, N/mm2

6

R = Inside radius, m

S = Stress value of material, N/mm2

E = Joint efficiency

t = Wall thickness

So the wall thickness is

ii) Longitudinal stress with condition t < R/2 or P < 1.25 SE

where


R = Inside radius, m

S = Stress value of material, N/mm2

E = Joint efficiency

So the wall thickness is

By comparing those two values, resulting maximum value of

reactor thickness. Hereby, minimum thickness of the reactor cylinder is, tmin=14.04

mm with 4 mm corrosion allowance because expecting the severe operating

conditions where erosion will occur, tdesign = 18.0 mm.

3.1.4.4 Domed head

(i) Try standard dished head (torisphere)

Crown radius Rc = Di = 4.80 m

Knuckle radius = 6 percent Rc = 0.288 m

A head of this size would be formed by pressing: no joints, so J=1

Where


D = Inside diameter, m

f = Stress value of material, N/mm2

7

= Stress concentration factor for torispherical heads

Therefore,

(ii) Try a “standard” ellipsoidal head, ratio major : minor axes = 2 : 1

Where


D = Inside diameter, m


J = Joint efficiency

Therefore,

So an ellipsoidal head would probably be the most economical. Take as

same thickness allowance of 4 mm as wall 18.0 mm.

Flat head

Where


De = Bolt diameter, m


Cp = Joint efficiency

Use bolted cover with a full face gasket Cp = 0.4

De= bolt circle diameter, take as approx. 4.80 m.

Therefore, from equation 5.46

This shows the inefficiency of a flat cover. It would be better to use a

flanged domed head. So, ellipsoidal head will be used as domed head for reactor.

8

3.1.4.5 Tube sheet (plate)

Tube sheet forms the barrier between shell and tube fluids, and where it is

essential for safety or process reason to prevent any possibility of intermixing due to

leakage at the tube sheet joint, double tube-sheets can be used, with the space

between the sheet vented. The thickness of tube sheet will reduce the effective

length of the tube slightly, and this should be allowed for when calculating the area

available for heat transfer. The thickness of tube sheet calculation given by the

TEMA standard as below

Thickness of tube sheet

Where

and

Where

= Outlet diameter of shell, mm

= Outlet diameter of tube, mm

= Number of tube

= Thickness of tube, mm

= Thickness of shell, mm

= Design pressure, N/mm2

= Design stress, N/mm2

= Elastic modulus of shell, N/mm2

= Elastic modulus of tube, N/mm2


Substituted k value into equation 5.48

Substituted F value inside equation 5.47

9

3.1.4.6 Reactor load

3.1.4.6.1 Weight of a cylindrical vessel with domed end

Where

Wv = total weight of the shell, excluding internal fittings, such as plates, N,

Cv = a factor to account for the weight of nozzles, man ways, internal supports,

etc; which can be taken as

= 1.08 for vessels with only a few internal fittings,

= 1.15 for distillation columns, or similar vessels, with several man ways,

and with plate support rings, or equivalent fittings,

Hv = height, or length, between tangent lines (the length of the cylindrical

section) = 15 m

g = gravitational acceleration, 9.81 m/s2,

t = wall thickness = 18.0 mm

pm = density of vessel material = 7787 kg/m3,

Dm = mean diameter of vessel D =4.818 m.

Cv taken is 1.08 for a few internal fittings.


3.1.4.6.2 Weight of tubes

Density of stainless steel 316 = 7787 kg/m3

(Obtain from Incropera De Witt, Heat and Mass Transfer)

From Pressure Vessel Handbook 10th Edition

For 2-in tube, 1 foot of pipe has weight 3.652 lb (Properties are based on ANSI B-

36.19)

0.3048 m = 1 ft

3.652 lbm=1.6565 kg

Therefore for 15m length = 81.5207 kg

Where

= Mass of single tubes, kg

10

= Gravitational force, m/s

= Number of tubes

Therefore,

3.1.4.6.3 Weight of fluid in the tube

Total volume of fluid inside tube

Where

di = Inside diameter of tubes, m

= Length of reactor, m

= Number of tubes inside reactor

Therefore from equation 5.49

Weight of fluid inside the tubes

Where

= Density of fluid, kg/m3

= Volume of tube, m3

= Gravitational force, 9.81 m/s2


3.1.4.6.4 Weight of tube sheet

Thickness of tube sheet = 25mm

Volume of tube sheet

Where

= Diameter of tube sheet = inside diameter of shell, m

= Length of tube sheet = tube sheet thickness, m

11


There are 2 tube sheet been used in the reactor. So, volume of tube sheet multiplied

by 2=0.9048 m3

Weight of tube sheet

Where

= Density of fluid, kg/m3

= Volume of tube, m3

= Gravitational force, 9.81 m/s2


Density of stainless steel 316 = 7787 kg/m3

3.1.4.6.5 Baffle weight

Volume of baffle, = 1.06 m3

3.1.4.6.6 Total weight

Therefore,

3.1.4.7 Analysis of shear stress and direct stress

3.1.4.7.1 Shear stress

i) Tangential stress

Where


= Inside diameter, mm



12

ii) Longitudinal stress

Where


= Inside diameter, mm



3.1.4.7.2 Direct stress

Direct stress is the stress that generated by the fluid inside vessel and its

vessel weight

Where

= Total weight of reactor (shell), kN

= Inside diameter, m

= Thickness of shell, m


3.1.4.8 Support

Support saddle used to support the container in a horizontal reactor. The

former is supported by two saddles can be considered as a simple supported beam

with uniformly distributed load. The distribution of the longitudinal axis of the bending

moment is shown in the diagram below:

13

The maximum point occurs on both sides and support the middle range. In

theory, the optimum support position, giving rise to the maximum bending moment is

the lowest position when the magnitude of the maximum value on both sides is

equal to the value of support in the middle of the range of:

1 22L LM M

Where

A = Distance from the tangent to the saddle support, m

L = Length of the container, the tangent line, m

H = column depth, m

= 1.218 m

Q = Total weight/saddle, N

= Total weight/2

= 1144.6171 kN

R = Radius of reactor

= 2.4 m

b = width of saddle, m

Bending moments at the two saddle supports, and bending in the middle of

the range, can be determined using the following equations:

14

Balance from the bending moment:

Solving from above equation, value for A =3.97m

Therefore

3.1.4.9 Stresses in vessel wall

Bending stress is a stress that cause by the bending moment in the shell

(vessel), bending moment is classified as the stress generated as a resultant to the

dead weight of reactor in horizontal position supported by the saddle support.

Bending stress longitudinal to the cross sectional area of shell as

Where

1LM = Longitudinal bending stress at mid-span

hI = Second moment of area of the shell

D = Shell diameter

t = Shell thickness

Therefore,

Resultant axial stress due to bending and pressure is given by:

Where

= Longitudinal bending moment at the support

= an empirical constant: 1

15

Downwind stress

Therefore,

Upwind stress

Therefore,

Principal stress,

Longitudinal stress,

The difference in principal stresses and the longitudinal stress resultant,

Because of the stress difference is <the maximum stress, S, the design is

acceptable.

The magnitude of the longitudinal bending stress on the strengthening of

support will depend on the local shell. If the shell does not remain round when

loaded, this means that some of the top cross section is not effective against

longitudinal bending. This stress is given as follows:

Where

= Longitudinal bending moment at the support

= an empirical constant: 1.0 for stiffened shell.

Therefore,

Because the value of b, 2 is smaller than the maximum design stress

allowable S, then the pressure vessel design of the heat exchanger is acceptable.

16

3.1.4.10 Saddle design

Saddle must be designed to withstand heavy loads caused by the container

and its contents. This saddle is made of stainless steel plate. Typically, the contact

angle cannot be less than 120° and not more than 1500. Smooth plates (wear plate)

are usually welded to the shell wall to reinforce the wall area in contact with the

saddle.

Saddle support design procedure given by Brownell and Young (1959) and

Megyesy (1977), the former equal to the diameter of 4.86 m, standard steel saddles

to container with a diameter of 4.8 m is used after interpolation been made as

shown in Table 5.3.

Table 5.3 Standard steel saddle

V

essel

Diameter

(m)

Dimension (m) mm

V Y C E J G

t

2

t

1

B

olt

diameter

B

olt

hole

4

.80

4

.303

0

.525

6

.99

3

.07

1

.852

0

.150

1

6

1

2

2

7

3

3

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3.1.4.11 Design bolt flange connection

Flange can be used in the body of the container when the container must

be divided into several sections for easy removal and maintenance. Flange

connection used to connect pipes to other equipment such as pumps and valves.

Typically used for connecting the connection of bolt with small diameter pipes, less

than 40 mm. Flange connections are also used to attach sections of pipe on the

installation and opening of facilities needed for maintenance, but the structure of the

pipe is usually welded to reduce costs.

Flange sizes vary, from a few millimeters in diameter for small pipes to

several meters in diameter for use as a body or head flange on the container. There

are four openings in the design of the reactor tube and shell, which requires the use

of connection, namely:

1. Welding-neck flanges.

2. Slip-on flanges, hub and plate types.

3. Lap-joint flanges.

4. Screwed flanges.

5. Blank, or blind, flanges.

Welded-neck flange type (steel) used for opening the input and output

openings for the connection and the nozzle of the reactor tube and shell. Given the

pressure vessel is operated under the operating pressure of 6.5 bar (650 kPa) at a

temperature of 155 C design, the flange of this type is selected for its ability to

withstand extreme operating conditions likely to be exposed to temperature loading,

shear, and vibration.

Optimum size for the flange to the nozzles feed (input) and the output of the

shell and tube can be determined using the following equation proposed by Sinnot:

18

Optimum pipe diameter at inlet stream from reactor 1

Data required:

G = 10313.2838 kg/hr= 2.8648 kg/s

= 7.471 kg/m3

Nom.

size

Pipe

o.d

d1

Flange Raised

face Bolting Drilling Neck

D b h1 d4 f No. d2 k d3 h2 r

200 219.1 320 20 55 258 3 M16 8 18 280 236 15 10

Optimum pipe diameter at inlet stream from splitter

Data required:

G = 7787.8489 kg/hr= 2.1633 kg/s

= 1301.2 kg/m3

Nom.

size

Pipe

o.d

d1

Flange Raised



25 33.7 100 14 35 60 2 M10 4 11 75 42 6 4

19

Optimum pipe diameter at outlet stream of reactor

Data required:

G = 18101.1257 kg/hr= 5.0281 kg/s

= 77.74 kg/m3

Nom.

size

Pipe

o.d

d1

Flange Raised



125 139.7 240 18 48 178 3 M16 8 18 200 155 10 8

Optimum pipe diameter at outlet from reactor for cooling system

Data required:

G = 5526.327 kg/s

= 1001.1462 kg/m3

Nom.

size

Pipe

o.d

d1

Flange Raised



200 219.1 320 20 55 258 3 M16 8 18 280 236 15 10

Pipe thickness. Equation below is follow British Standard 5500.

Where

P : Internal pressure,bar

D : Pipe outer diameter, mm

: Design stress at working temperature, N/mm2

Inlet from reactor 1

From equation 5.65, thickness of nozzle is

20

Inlet from splitter


Outlet from reactor


Optimum pipe diameter at outlet from reactor for cooling system


Plug flow reactor data

sheet

Equipment no.:Plug flow reactor (PFR-101)

Description : Convert ammonia and nitric acid to

ammonium nitrate

Sheet no:

Operating Data

No.

Required. 1 Capacity 109.87 m

3

Specific

gravity of

content

0.107 Computed (yes or no)

Shell

Content Ammonia, water, nitric acid and ammonium nitrate

Length 15 m

Max.

working

pressure

133.5 N/mm2

Design

Pressure 0.778 N/mm

2

21

Working

temp. 458.15 K

Design

temp. 203.5

0C

Material Stainless steel 16Cr-2Mo-8Ni (316)

Joint factor 1.0

Corrosion

allowance 4 mm

Shell

thickness 18.00 mm

Type of

head

Elipsoid

al Thickness 18.00 mm

Reactor load 2289.2342 kN

Tangetial

stress 51.8667 N/mm

2

Longitudina

l stress 103.733 N/mm

2

Direct stress 8402.3555 N/mm2

Type of

support Saddle

Distance of

tangent to

saddle

support

3.97 m

Tube

Tube outside

diameter 2.735 in

Tube inside

diameter 2.067 in

Wall

thickness 0.154 in

Number of

tube 1933

22

required

Area of tube 0.003790 m2

Volumetric

flow rate 0.08776 m

3/hr

Bundle

diameter 4.30 m

Shell inside

diameter-

diameter-

bundle

diameter

0.50 m

Shell

diameter 4.80 m

Number of

baffle 8

Distance

between

baffle

1.92 m

Pitch

diameter 0.0868 m

Tube sheet

thickness 0.1795 m

Cooling system

Fluid Water

Velocity of

fluid 3 m/s

Flow rate 5526.327 kg/hr

Fluid inlet

temperature 25

0C

Fluid outlet

temperature 81.70

0C

Tube side

coefficient 3715.3883 W/m

2.K

23

Shell side

coefficient 14885.00 W/m

2.K

Tube side

pressure

drop

0.46 bar

Shell side

pressure

drop

8.458 bar

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3.2 REACTOR 2, R2

3.2.1 Design Pressure

A vessel must be designed to withstand the maximum pressure to which it

is likely to be subjected in operation. For vessels under internal pressure, the design

pressure is normally taken as the pressure at which the relief device is set. This will

normally be 5 to 10 per cent above the normal working pressure, to avoid spurious

operation during minor process upsets. In this design, considering 10 % safety

factor so that the design pressure become as below:

(1.36)

3.2.2 Design Temperature

The operating temperature of our reactor is taken as 185 0C. For safety

reason, the design pressure of this reactor is taken as 10% above the operating

temperature to avoid spurious operation during minor process upsets.

(1.37)

0C

K

3.2.3 Material Of Construction

Many factors have to be considered when selecting engineering materials

but for chemical process plant the overriding consideration is usually the ability to

resist corrosion. The material selected must have sufficient strength and be easily

worked. The most economical material that satisfies both process and mechanical

requirements should be selected which is this will be the material that gives the

lowest cost over the working life of the plant and allowing for maintenance and

replacement.

Stainless steels are the most frequently used corrosion resistant materials in

the chemical industry. To impart corrosion resistance the chromium content must be

above 12 per cent and the higher the chromium content, the more resistant is the

alloy to corrosion in oxidising conditions. Nickel is added to improve the corrosion

resistance in non-oxidising environments.

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A wide range of stainless steels is available, with compositions tailored to

give the properties required for specific applications. Type 304 also-called 18/8

stainless steels is the most generally used stainless steel. It contains the minimum

Cr and Ni that give a stable austenitic structure. The carbon content is low enough

for heat treatment not to be normally needed with thin sections to prevent weld

decay. The uniform structure of austenitic is the structure desired for corrosion

resistance and it is these grades that are widely used in the chemical industry. The

austenitic stainless steels have greater strength than the plain carbon steels,

particularly at elevated temperatures (see Appendix A1). So, as conclusion stainless

steels type 304 is the best material of construction and then selected as material of

construction for the reactor.

3.2.4 Determination Of Minimum Thickness Of The Reactor

(1.38)

Where:

, minimum thickness

Pi , the design pressure

Di , the inside diameter

f, design stress

The strength of metals decreases with increasing temperature, so the

maximum allowable design stress will depend on the material temperature. The

design temperature at which the design stress is evaluated should be taken as the

maximum working temperature of the material. With design temperature is equal to

maximum operating temperature, 185 oC, design stress for stainless steel 304, is f =

115 N/mm2 = 115 bar (R.K. Sinnot, 1999. Chemical Engineering Design). Typical

design stress values for some common materials are shown in Appendix A2.

Thus from Eqn. (1.38),

26

The corrosion allowance is the additional thickness of metal added to allow

for material lost by corrosion and erosion, or scaling. Corrosion is a complex

phenomenon and it is not possible to give specific rules for the estimation of the

corrosion allowance required for all circumstances. The allowance should be based

on experience with the material of construction under similar service conditions to

those for the proposed design. For carbon and low-alloy steels, where severe

corrosion is not expected, a minimum allowance of 2.0 mm should be used.

Add allowance for corrosion = + 0.002 m =

3.2.5 Design of Vessel Heads

The end of a cylindrical vessel is closed by heads of various shapes. The

common types used are:

i. Flat heads

ii. Hemispherical heads

iii. Ellipsoidal heads

iv. Torispherical heads

The heads used for the vessel may be flat if they are suitably buttressed

but preferably they are some curved shape as the hemispherical, ellipsoidal or

torispherical heads. Standard torispherical heads (dished ends) are the most

commonly used end closure for vessels up to operating pressures of 15 bar. They

can be used for higher pressures, but above 10 bar their cost should be compared

with that of an equivalent ellipsoidal head. Above 15 bar an ellipsoidal head will

usually prove to be the most economical closure to use.

The minimum thickness of torispherical and ellipsoidal head can be

calculated by using equation below:

For torispherical heads,

(1.39)

Where

Pi , internal pressure

J , joint factor =1

f, design stress

Rc, crown radius = Di

27

Cs, stress concentration factor = ¼(3+( Rc/Rk)1/2)

Rk, knuckle radius =0.06 Rc

From earlier calculation,

Pi = 8.8 bar

Rc = Di = 4.3151 m (1.40)

Rk =0.06 Rc = 0.06(4.3151) (1.41)

= 0.2589 m

Cs = ¼(3 + ( Rc/Rk)1/2) (1.42)

= ¼ (3 + (4.3151/0.2589)1/2)

= 1.7706

f = 115 N/mm2 = 115 bar

From Eqn. (1.39),

e =

= 0.0015 m

For ellipsoidal heads,

(1.43)

Where

Pi , internal pressure

J , joint factor =1

f, design stress

Di, inside diameter

From Eqn. (1.25),

e =

= 0.1664 m

By comparing minimum thickness between torispherical and ellipsoidal head,

torispherical head is the most economical. So, torispherical head is choosen for the

design domed heads. Hence,

Add 0.002 m allowance for corrosion = 0.0015 + 0.002 m = 0.0035 m

i

ii

PJf

DPe

2.02

28

3.2.6 Determination of Piping Sizing

Liquids particularly can be transported through pipelines with pumps,

blowers, compressors or ejectors. Standard pipe is made in a discrete number of

sizes that are designed by nominal diameters (R.K. Sinnot, 1999. Chemical

Engineering Design).

Formula for Optimum diameter for stainless steel pipe is as follow:

37.052.0260, Goptimumd (1.44)

Where:

G = mass flow, kg/s

= density of flow, kg/m3

3.2.7.1 Diameter pipe for flow in

mm

optimumd

0124.97

)7117.67()0146.3(260, 37.052.0

From appendix A3 nominal diameter d = 80 mm.

3.2.7.2 Diameter pipe for flow out

mm

optimumd

5182.216

)4710.7()9417.2(260, 37.052.0

From appendix A3 nominal diameter d = 200 mm.

3.2.6.3 Thickness of nozzle pipe inlet

Calculation of thickness of nozzle pipe inlet is as follow:

t =

(1.45)

Where:

P = internal pressure, bar

d = pipe od, mm

σd = design stress at working temperature, N/mm2

From Eqn. (1.45),

29

t =

= 0.3049 mm

Add 2 mm allowance for corrosion = 0.3049 + 2 = 2.3049 mm.

3.2.6.4 Thickness of nozzle pipe outlet

Calculation of thickness of nozzle pipe outlet is as follow:

t =

(1.46)

Where:

P = internal pressure, bar

d = pipe od, mm

σd = design stress at working temperature, N/mm2

From Eqn. (1.46),

t =

= 0.7623 mm

Add 2 mm allowance for corrosion = 0. 7623 + 2 = 2. 7623 mm.

3.2.7 Design Of Reactor Vessel Subject To Combined Loading

Pressure vessels are subjected to other loads in addition to pressure and

must be designed to withstand the worst combination of loading without failure. The

main sources of load to consider are:

i. Pressure

ii. Dead weight of vessel and contents

iii. External loads imposed by piping and attached equipments

3.2.7.1 Weight Loads

The major sources of dead weight loads are:

i. The vessel shell.

ii. The vessel fittings: manways, nozzles.

iii. Internal fittings: tubes, plates (plus the fluid on the plates), heating

and cooling coils.

30

iv. External fittings: ladders, platforms, piping.

v. The weight of liquid to fill the vessel. The vessel will be filled with

water for the hydraulic pressure test and may fill with process liquid

due to misoperation.

For preliminary calculations the approximate weight of a cylindrical vessel

with domed ends and uniform wall thickness, can be estimated from the following

equation:

(1.44)

Where,

Cv = a factor account for the weight of nozzles, man ways

= 1.08 for vessel with only a few internal fittings

= 1.15 for vessel with several man ways and other fittings

Dm = mean diameter of the vessel

= Dm + t

= 4.3151 m + 2(0.1737) m

= 4.6625 m

Hv = height/length of the cylindrical area

= 12.9453 m

Thus,

3.2.8 Vessel Support

The method used to support a vessel will depend on the size, shape and

weight of the vessel, the design temperature and pressure, the vessel location and

arrangement and the internal and external fittings and attachments. Horizontal

vessels are usually mounted on two saddle supports (see Appendix A4). The

supports must be designed to carry the weight of the vessel and contents, and any

superimposed loads, such as wind loads. Supports will impose localized loads on

the vessel wall and the design must be checked to ensure that the resulting stress

concentrations are below the maximum allowable design stress. Supports should be

31

designed to allow easy access to the vessel and fittings for inspection and

maintenance.

Though saddles are the most commonly used support for horizontal

cylindrical vessels, legs can be used for small vessels. A horizontal vessel will

normally be supported at two cross-sections. If more than two saddles are used the

distribution of the loading is uncertain. For a uniformly loaded beam the position will

be at 21 per cent of the span, in from each end. The saddle supports for a vessel

will usually be located nearer the ends than this value to make use of the stiffening

effect of the ends.

The saddles must be designed to withstand the load imposed by the weight

of the vessel and contents. They are constructed of bricks or concrete or are

fabricated from steel plate. The contact angle should not be less than 120o and will

not normally be greater than 150o. Wear plates are often welded to the shell wall to

reinforce the wall over the area of contact with the saddle. The dimensions of typical

standard saddle designs are given in figure below:

Figure 1.4: The Dimensions of Typical Standard Saddle Designs

(Source: Sinnott, R.K, 1999. Coulson & Richardson’s Chemical

Engineering, Vol. 6: “Chemical Engineering Design”, Oxford, Butterworth-

Heinemann).

32

3.2.9 Type Of Flange And Selection

Flanged joints are used for connecting pipes and instruments to vessels, for

manhole covers and for removable vessel heads when ease of access is required.

Flanges may also be used on the vessel body when it is necessary to divide the

vessel into sections for transport or maintenance. Flanged joints are also used to

connect pipes to other equipment such as pumps and valves. Screwed joints are

often used for small-diameter pipe connections below 40 mm.

Several different types of flange are used for various applications. The

principal types used in the process industries are:

i. Welding-neck flanges

ii. Slip-on flanges, hub and plate types

iii. Lap-joint flanges

iv. Screwed flanges

v. Blank or blind, flanges

Welding-neck flanges (see Appendix A5 (a)) have a long tapered hub

between the flange ring and the welded joint. This gradual transition of the section

reduces the discontinuity stresses between the flange and branch, and increases

the strength of the flange assembly. Welding-neck flanges are suitable for extreme

service conditions where the flange is likely to be subjected to temperature, shear

and vibration loads. They will normally be specified for the connections and nozzles

on process vessels and process equipment.

Slip-on flanges (see Appendix A5 (b)) slip over the pipe or nozzle and are

welded externally and usually also internally. The end of the pipe is set back from 0

to 2.0 mm. The strength of a slip-on flange is from one-third to two-thirds that of the

corresponding standard welding-neck flange. Slip-on flanges are cheaper than

welding-neck flanges and are easier to align but have poor resistance to shock and

vibration loads. Slip-on flanges are generally used for pipe work.

Lap-joint flanges (see Appendix A5 (c)) are used for piped work and most

suitable in this design reactor. They are economical when used with expensive alloy

pipe such as stainless steel as the flange can be made from inexpensive carbon

steel. Usually a short lapped nozzle is welded to the pipe but with some schedules

of pipe the lap can be formed on the pipe itself and this will give a cheap method of

pipe assembly.

Screwed flanges (see Appendix A5 (d)) are used to connect screwed

fittings to flanges. They are also sometimes used for alloy pipe which is difficult to

weld satisfactorily. Blind flanges (blank flanges) are flat plates, used to blank off

33

flange connections, and as covers for manholes and inspection ports. So, in this

design lap joint flange is chosen as the best flange.

3.2.10 Gasket

Gaskets are used to make a leak-tight joint between two surfaces. It is

impractical to machine flanges to the degree of surface finish that would be required

to make a satisfactory seal under pressure without a gasket. The following factors

must be considered when selecting a gasket material:

i. The process conditions: pressure, temperature, corrosive nature of the

process fluid.

ii. Whether repeated assembly and disassembly of the joint is required.

iii. The type of flange and flange face

Up to pressures of 20 bar, the operating temperature and corrosiveness of the

process fluid will be the controlling factor in gasket selection. Vegetable fibre and

synthetic rubber gaskets can be used at temperatures of up to 100 oC. Solid

polyfluorocarbon (Teflon) and compressed asbestos gaskets can be used to a

maximum temperature of about 260 oC. Metal-reinforced gaskets can be used up to

around 450 oC. Plain soft metal gaskets are normally used for higher temperatures.

So, compressed asbestos is chosen as the best gasket to be used in this reactor

design (see Appendix A6).

3.2.11 Flange Faces

Flanges are also classified according to the type of flange face used. There

are two basic types:

i. Full-faced flanges (see Appendix A7 (a)) where the face contact area

extends outside the circle of bolts; over the full face of the flange.

ii. Narrow-faced flanges (see Appendix A7 (b,c,d) where the face

contact area is located within the circle of bolts.

Full face, wide-faced, flanges are simple and inexpensive but are only

suitable for low pressures. The gasket area is large and an excessively high bolt

tension would be needed to achieve sufficient gasket pressure to maintain a good

seal at high operating pressures. The raised face, narrow-faced, flange shown in

34

Appendix A7 (b) is probably the most commonly used type of flange for process

equipment.

Where the flange has a plain face, as in Appendix A7 (b), the gasket is held

in place by friction between the gasket and flange surface. In the spigot and socket,

and tongue and grooved faces, Appendix A7 (c), the gasket is confined in a groove

which prevents failure by blow-out. Matched pairs of flanges are required, which

increases the cost, but this type is suitable for high pressure and high vacuum

service. Ring joint flanges, Appendix A7 (d), are used for high temperatures and

high pressure services. So, in this design raised face, narrow-faced is chosen as the

best flange faces.

3.2.11 CONCLUSION

In this work, the design of plug flow reactor has successfully been carried

out. From the calculation, the volume of the vessel is 189.3128 m3 with 4.3151

diameter and 12.9453 length. The detail information of the design is as presented in

Table 1.2 and Table 1.3.

Table 1.3: Summary of mechanical design

Operating pressure, bar 8.8

Operating temperature, k 476.65

Thickness of reactor, m 0.1737

Type of head Torispherical

Total weight of reactor, N 3500

Vessel support Saddle support

Type of flanges Lap-joint flange

Gasket Compressed asbestos

Flange Faces Raised face, narrow-faced

35

3.3 FALLING FILM EVAPORATOR 1, F1

In designing a chemical plant, the mechanical design of the process equipments

such as pressure vessel, heat exchanger tube sheets, storage tanks, centrifuges

and etc are needed. The detailed mechanical designing of equipment is done by

mechanical engineers who are more familiar with the codes and design. On the

other hand, chemical engineer will be responsible in developing and specifying the

basic design information for particular equipment for specialist designer.

For falling-film evaporator, the data for mechanical design needed are:

i. Vessel function

ii. Process materials and services

iii. Operating and design temperature and pressure

iv. Materials of construction

v. Vessel dimensions and orientation

vi. Types of vessel heads to be used

vii. Openings and connections required

viii. Specification of internal fitting

3.3.1 Design Pressure

In designing a vessel, it needs to withstand the maximum pressure during

operation. For a vessel that is subjected to vacuum, the design should resist

the maximum differential pressure and is designed for full negative pressure

of 1 bar, unless it is fitted with an effective vacuum breaker.

The design pressure should be taken to be 10% above the normal

operating pressure:

3.3.2 Design Temperature

Since the strength of metals decreases with increasing temperature, the

maximum allowable design stress is evaluated at design temperature which

is the maximum working temperature of the material.

The design temperature can be evaluated with 5% safety factor above the

operating temperature:

36

3.3.3 Materials of Construction

Typically, the pressure vessel is made of plain carbon steel, low and high

alloy steels, alloys and etc. The material is selected based on its suitability

with the process environment and fabrication.

For the falling-film evaporator, the shell are filled with hot steam, thus,

constructed with stainless steel (SS304) while the tubes are constructed

from stainless steel (SS316) due to the mild corrosive of the feed which is

the ammonium nitrate solution of 72 wt%.

3.3.4 Design Stress

For the purpose of design, the value of maximum allowable stress that can

be accepted in the material of construction is needed. For the material to

able to withstand without failure under standard condition, a suitable design

stress factor (factor of safety) is applied to the maximum stress of the

material. This design stress factor is to cover any uncertainties in the design

methods, the loading, the quality of materials, and the workmanship. The

value can be taken from Appendix B.1 and typical design stress for material

can be taken from Appendix B.2.

3.3.5 Welded Joint Efficiency, and Construction Categories

The welded joint strength depends on the type of joint and the quality of the

welding. The allowable design stress of the material multiplied by a welded

joint factor will give the possible lower strength of a welded joint compared to

a virgin plate. Typical value of J is given in Appendix B.3. For the design of

this evaporator, J of 1.0 is taken because this value means that the joint is

equally strong as the virgin plate.

3.3.6 Corrosion Allowance

Corrosion allowance is the additional thickness of the metal to the design to

allow for corrosion and erosion, or scaling. The corrosion allowance for this

evaporator is 4mm because, the process material used in this equipment, i.e.

ammonium nitrate solution (75wt %-84wt %) may cause corrosion and

scaling to the equipment.

37

3.3.7 Design Loads

This equipment should be designed to resist loading at which a pressure

vessel will be subject during service. It can be divided into major and

subsidiary loads. Major load includes design pressure, maximum weight of

vessel and contents at operating temperature and hydraulic test condition,

wind loads, loads supported or reacting on the vessel. Subsidiary loads

includes local stresses caused by supports, internal structures and

connecting pipes; shock loads, bending moments, stresses due to difference

in temperature and loads caused by fluctuations in temperature and

pressure. Design load is further discussed in Section 2.4.

3.3.8 Minimum Practical Wall Thickness

The wall thickness should not be less than the value given below. (Include

corrosion allowance of 2mm)

Figure 2.1: Minimum practical wall thickness

3.3.9 Cylindrical Shells

The minimum thickness required to resist internal pressure is given by:

Where:

38

Process vessels that are operated under vacuum are subjected to external

pressure. The maximum pressure it will subject to is 1 bar (1 atm). In determining

the wall thickness required for process vessel subjected to external pressure, it is

required to know the failure through elastic instability (buckling).

The critical pressure to cause buckling, PC for long vessel with stiffening

ring is given by:

, value from Appendix B.4

3.3.10 Design of Stiffness Rings

Figure 2.2: Stiffness Ring

Load per unit length,

Second moment of area of the ring to avoid buckling,

Factor of safety taken as 6,

Critical load to cause buckling in a ring under uniform radial load, :

39

3.3.11 Vessel Head

Vessel head are used as a closure of a cylindrical vessel.

Figure 2.1: Typical Head and Closure

3.3.11.1 Torispherical heads

For vessel subjected to internal pressure, the minimum thickness of

torispherical head is:

Where:

To avoid buckling, the ratio of knuckle to crown radii should not be

less than 0.06, and the crown radius should not be greater than the

diameter of the cylindrical section.

When it is subjected to external pressure,

Minimum vessel thickness,

(f)

(g)

(h)

40

For torispherical, radius Rs is equivalent to Crown radius, Rc

3.3.11.2 Ellipsoidal heads

For vessel subjected to internal pressure, the minimum thickness of

ellipsoidal head is:

When subjected to external pressure,

Minimum vessel thickness,

For ellipsoidal,

,

Where 2a = major axis = Do,

2b = minor axis = 2h,

h = height of the head from the tangent line.

3.3.11.3 Flat ends

Minimum thickness of flat end required for internal pressure:

Where:

For typical design, the design constant and nominal diameter area as

follows:

From Figure 2.1,

i. (a) is flanged plate, for diameters less than 0.6m and corner radii

at least equal to 0.25e (Cp=0.45, De=Di);

ii. (b) and (c) is welded plate where the plate is welded to the end of

the shell with a fillet weld with angle of fillet of 45 and depth

equal to the plate thickness (Cp=0.55,De=Di)

iii. (d) is bolted cover with full gasket (Cp=0.4,De=bolt circle

diameter)

iv. (e) is bolted end-cover with a narrow-face gasket

(Cp=0.55,De=mean diameter of gasket)

41

3.3.12 Stresses Analysis

Primary Stresses:

Longitudinal and circumferential stresses due to internal or external

pressure:

Direct stress weight,

The dead weight stress will be tensile (positive) for points below the

plane of vessel supports, and compressive (negative) for points above

the supports.

Bending stress,

Where:

Torsional shear stresses,

This stress is resulted from torque caused by loads offset from the vessel

axis. This load is usually small and need not be considered in preliminary

design.

Principal Stresses:

Where:

Total longitudinal stress,

If torsional shear stress, is negligible, principal stress will be

42

Compressive stress and elastic stability:

If the resultant axial stress, due to the combined loading is compressive,

the failure of the vessel may be due to elastic instability (buckling). The

design must be check to make sure that the maximum value of the resultant

axial stress does not exceed the critical value at which buckling will occur.

Critical buckling stress,

3.3.13 Weight Loads

The weight loads comprises of:

i. Vessel Shell

The approximate weight of a cylindrical vessel with domed ends, and uniform

wall thickness,

Weight of Vessel:

Where:

ii. Vessel Fittings

For vessel fittings, the following can be used:

(a) Caged ladders, steel, 360

length

(b) Plain ladders, steel, 150

length

(c) Platforms, steel, for vertical columns, 1.7

area

(d) Contacting plates, steel including typical liquid loading, 1.2

plate

area

43

For Internal Fittings, i.e. tubes:

Weight of Tubes:

Where:

iii. Wind Loads

For tall columns installed in the open, it is important to consider wind loading.

A wind speed of 160 km/h is usually taken for preliminary design which is

equivalent to 1280

wind pressure. The wind velocity is lower near the

ground than higher ground.

For a smooth cylindrical column or stack,

Dynamic wind pressure:

wind velocity, km/h

The loading per unit length of the column:

For a uniformly loaded cantilever the bending moment at any plane:

44

3.3.14 Skirt Supports

The skirt carried the load and is transmit to the foundation slab by the skirt

base ring (bearing plate). The moment produced by wind and other lateral

loads will tend to overturn the vessel. This will be opposed by the couple set

up by the weight of the vessel and the tensile load in the anchor bolts. Many

types of base ring designs as shown in Figure 2.1 is used with skirt support,

for example, rolled angle and plain flange rings suitable for small vessel and

double ring stiffened by gussets.

Figure 2.1: Flange ring design

Base Ring and Anchor Bolts:

The carried load by the skirt is transferred to the base ring or the foundation

slab (bearing plate). Winds and other loads produces moment that will tend

to overturn the vessel. The couple set up by the weight of the vessel and the

tensile load in the anchor bolt in turn, will oppose to the moment.

The following is the guide rules when selecting the anchor bolts given by

Scheiman:

Bolts smaller than 25mm diameter should not be used

Minimum number of bolts is 8

Use multiple number of 4 bolts

Bolt pitch should not be less than 600 mm

Approximate pitch circle diameter

Circumference of bolt circle

Minimum recommended bolt spacing

Number of bolts required, at minimum recommended bolt spacing

45

Assuming the anchor bolts share the overturning load equally,

Bolt area required,

Where:

Bolt root diameter

Total compressive load on the base ring per unit length,

Taking the bearing pressure, as 5

Minimum width of the base ring,

Choose suitable anchor bolt size design from Appendix ???.

Actual width required

Actual bearing pressure on concrete foundation:

Minimum thickness for the base ring,

Skirt Thickness:

By trial and error, choose

The maximum dead weight load on the skirt occurs when the vessel is full

with water.

Use data acquired previously,

46

Total weight of skirt

Wind loading,

Bending moment at base of skirt,

By trial and error,

Assume skirt thickness,

Previously,

Bending stress in the skirt,

Dead weight stress in the skirt,

At test condition, the vessel full of water for the hydraulic test,

,

At operating condition,

Maximum

Maximum

Take joint factor,

(Double-welded butt or equivalent type of joint and degree of radiography is

spot)

Criteria for design:

Maximum

Maximum

Both criteria are satisfied, add 4 mm for corrosion.

47

3.3.15 Piping and Flanges

Optimum diameter of flange:

Where:

Nozzle thickness:

Where:

3.3.16 Evaporator Tube-Plates

Tube-plates support the tubes, and separate the shell and tube side fluids.

Since, one side is subjected to shell-side pressure and tube-side pressure on

the other side. Therefore, the design must able to support the maximum

differential pressure that is likely to occur.

A tube plate is a perforated plate with an unperforated rim, supported at its

periphery. The holes of plate for the tubes weaken the plate and reduce its

flexural rigidity. In between the holes is a material that holds the holes

together is ligament. The presence of tubes strengthens the plate.

Ligament efficiency of perforated plate,

Where:

The plate must be thick enough to resist the bending and shear stresses

caused by the pressure load and any differential expansion of the shell and

tube.

48

The minimum plate thickness to resist bending can be estimated by:

Where:

The value of is relies on the type of head,

Shear stress in the tube plate can be calculated by equating the pressure

force on the plate to the shear force in the material at the plate periphery.

Minimum plate thickness to resist shear is given by:

The design thickness is taken as the greater of the values obtained from

bending and shears resistance and must be greater than the minimum

thickness given from Appendix B.5

49

3.3.17 Calculations

Design Pressure, PD and External Pressure, Pe:

Maximum pressure for vessel under external pressure is 1 bar,

At Tube-Side:

At Shell-Side:

Design Temperature, TD:

At Tube-Side:

At Shell-Side:

Design Stress (Nominal Design Stress):

Refer to Appendix B.1,

Shell Side:

Material of Construction : Stainless Steel (SS 304)

Typical Design Stress, f : 125.5 N/mm2 (calculated at T=165oC)

Tensile Strength : 510 N/mm2

Tube Side:

Material of Construction : Stainless Steel (SS 316)

Typical Design Stress, f : 143.55 N/mm2 (calculated at T=121.5oC)

Tensile Strength : 520 N/mm2

From Appendix B.1, design factor taken for Austenitic stainless steel

at minimum yield stress is 1.5. The design stresses for tubes and shells are

calculated from Appendix B.2 are 143.5

and 125.5

respectively.

Thus,

50

:

Welded joint efficiency, J and construction categories:


Welded joint factor chosen, J = 1

Type of joint:

Double-welded butt or equivalent of 100% degree of radiography.

Corrosion Allowance:

Since, moderate corrosions are expected in the tubes and shell, the

corrosion allowance of 4.0mm is used.

Design of Cylindrical Shells under Internal Pressure

Minimum thickness, e plus corrosion allowance of 4 mm =

Critical Pressure to Cause Buckling, PC:

For long vessel with stiffening ring, the critical pressure of buckling is high,


As

For this particular thickness, e = 4.0584mm, the design pressure is below of

critical pressure (

), thus the thickness is suitable

51

Design of Stiffness Ring:

Assume,

Load per unit length,

Second moment of area of the ring to avoid buckling,

Taken factor of safety = 6,

Critical load to cause buckling in a ring under uniform radial load, :

Since,

The length and diameter of stiffening ring are acceptable.

Vessel heads:

If using torispherical head,

Subjected to internal pressure

Where:

Plus corrosion allowance of 4mm,

52

Subjected to external pressure

For ammonium nitrate solution, corrosion allowance is 4 mm.

If using ellipsoidal head,

Subjected to internal pressure

Plus corrosion allowance,

Subjected to external pressure

=

For ammonium nitrate solution, corrosion allowance is 4 mm, thus

For flat ends with bolted cover with full gasket,

Take

Add corrosion allowance,

53

Design of Vessel Subject to Combined Loading

i. Weight Loads:

Weight of Vessel:

Weight of Tubes:

Weight of External Fittings:

Installed caged ladder, steel to the equipment,

Thus,

ii. Wind Loading:

Take wind velocity,

The load due to wind of smooth cylindrical column,

Since no thermal insulation and attachment,

Loading per unit length of column,

54

Bending moment at bottom tangent line,

iii. Analysis of Stresses

At bottom tangent line,

Pressure Stresses:

Dead weight stress:

The dead weight stress will be tensile (positive) for points below the plane of

vessel supports, and compressive (negative) for points above the supports.

Since calculated for points above the supports, it is compressive

(negative).

Bending stresses:

Bending stress will be compressive or tensile,

Where:

Resultant longitudinal stress:

Previously,

,

,

is compressive

(negative),

55

Since the torsional shear stress is negligible, the principle stress will be

and .

The radial stress is negligible,

Up-wind

12.7193

7.3349

Down-wind

6.5768

7.3349

The greatest difference between the principal stresses will be on the down-

wind side,

,

where it is well below the maximum allowable design stress of 125.5

.

iv. Elastic Stability (Buckling)

Previously, the resultant axial stress, due to the combined loading is

compressive, the failure of the vessel may be due to elastic instability

(buckling). The design must be check to make sure that the maximum value

of the resultant axial stress does not exceed the critical value at which

buckling will occur.

Critical buckling

stress,

The maximum compressive stress will occur when the vessel is not under

pressure

=

is well below the critical buckling

stress.

So the design is satisfactory.

56

v. Vessel Support: Skirt Support

For tall vertical vessels, skirt supports are preferred because they do not

lead to concentrated local loads on the shell, it offers less restraint against

differential thermal expansion, and reduce the effect of discontinuity stresses

at the junction of the cylindrical shell and the bottom. The skirt support shall

be provided with at least one opening for inspection.

Skirt thickness:

Try straight cylindrical skirt,

Material of Construction = Plain Carbon Steel

Design stress, f at ambient temperature =

Young’s Modulus at ambient temperature,

Height of Skirt,

The maximum dead weight load on the skirt occurs when the vessel is full

with water.

Previously,

Weight of vessel, ,

Total weight of skirt

Wind loading,

Bending moment at base of skirt,

By trial and error,

Take skirt thickness,

Previously,

Bending stress in the skirt,

Dead weight stress in the skirt,

At test condition, the vessel full of water for the hydraulic test,

,

57

At operating condition,

Maximum

Maximum

Take joint factor,

(Double-welded butt or equivalent type of joint and degree of radiography is

spot)

Criteria for design:

Maximum

Maximum

Both criteria are satisfied, add 2 mm for corrosion, which gives:

vi. Base Ring and Anchor Bolts

Approximate pitch circle diameter

Circumference of bolt circle

Minimum recommended bolt spacing

Number of bolts required, at minimum recommended bolt spacing

Bolt design stress,

(typical design value)

Take

Bolt area required,

Bolt root diameter

58


Taking the bearing pressure, as 5

Minimum width of the base ring,

Use M24 bolts (BS 4190:1967);

Nominal Diameter = 24 mm,

Root area = 353 ,

This is the minimum width required; actual width will depend on the chair

design.

Actual width required


Minimum thickness for the base ring,

Skirt to be welded flush with outer diameter of column shell.

vii. Tube-plates

Ligament efficiency of perforated plate,

The minimum plate thickness to resist bending can be estimated by:

Minimum plate thickness to resist shear is given by:

59

The design thickness is taken as the greater of the values obtained from

bending and shears resistance and must be greater than the minimum

thickness given from Appendix B.5

viii. Opening and Nozzles:


Nozzle thickness:

Feed Inlet:

Concentrate Outlet:

Vapor Outlet:

Steam Inlet:

60

Condensate Outlet:

3.3.18 SUMMARY

General Option

Identifier Heat Exchanger

Description Fixed Tube-Sheets, One pass shell

Shell Material Stainless Steel 304L

Tube Option

Tube Material Stainless Steel 316L

Tube Dimensions ,

Channel and Shell Option

Shell Material Stainless Steel 304L

Shell Dimension ,

Top-channel dimensions Type: Bonnet

Head: Ellipsoidal head

Bottom-channel dimensions Type: Bonnet

Head: Ellipsoidal head

Tube-sheet Options

Tube Layout Tube Count: 185 Tube Pitch: 47.625mm

Pattern: Equilateral Triangular

Tube-sheet Dimensions (top and bottom)

Material: Stainless Steel 304L

Thickness:

61

Design Conditions Summary

Design Conditions

Tube Side

Design Pressure

Design Temperature

Mean Temperature

Shell Side

Design Pressure

Design Temperature

Mean Temperature

Tube-sheet

Design Temperature

Vessel Support

Type Straight cylindrical skirt

Thickness

Material Plain Carbon Steel

Height 3m

Base Ring and Anchor Bolts

Number of Bolts Required 12

Bolts Nominal Diameter

Root Area

M24 24mm

353

Minimum Width of Base Ring 138 mm

Minimum Thickness of Base Ring 6 mm

Tube-plates

Diameter 753.1735 mm

Minimum Plate Thickness 29.9798 mm

Openings and Nozzles

Feed Inlet

Concentrate Outlet

Vapour Outlet

Steam Inlet

Condensate Outlet

Stresses Analysis

Weight Loads

Wind Loading

Dead Weight

Bending Stress

Elastic Stability

62

3.4 HEAT EXCHANGER

Shell side details :

o Material = carbon steel

o Number of shell passes = 1

o Working pressure = 0.8 N/mm2

o Design stress for carbon steel, J = 109 N/mm2

o Inlet temperature = 180 oC

o Outlet temperature = 104.1 oC

Tube side details :

o Number of tubes = 128

o Number of passes = 1

o Outside diameter = 19.5 mm

o Inside diameter = 16.5 mm

o Length = 5 m

o Pitch rectangular = 24.38 mm

o Inlet temperature = -40 oC

o Outlet temperature = 65 oC

3.5.1 Design pressure

The design pressure, normally taken 10% above the normal working

pressure

Design pressure, Pi = 1.1xPo

= 1.18.0 x

= 2/88.0 mmN

3.5.2 Design temperature

For the shell side and tube side, the highest operating temperatures are at

180oC, and add up 2oC for uncertainties in temperature prediction.

Design Temperature, Ti = CC oo 2180

= Co182

63

3.5.3 Material selection

Carbon steel is chosen because this material mostly used in industry and

the prices is cheapest. Besides, it is routinely used for most organic chemicals and

neutral or basic aqueous solutions at moderate temperatures.

From Table 13.2 page 812 Chemical Engineering Volume 6, the design

stress was obtain at operating temperature (T = 180 oC)

Design stress, 2/109 mmNf s

3.5.4 Welded joint efficiency

Joint efficiency was selected to be 1.0 because this implies that the joint is

equally as strong the virgin plate, complete weld length, and remaking any defects.

The lower joint factor will result in a thicker and heavier vessel.

Welded joint efficiency, 0.1J

3.5.5 Corrosion allowance

The corrosion allowance is the additional thickness of metal added to allow

for material lost by corrosion and erosion, or scaling. For carbon steel, where sever

corrosion is not expected, a minimum allowance of 2.0 mm should be used.

3.5.6 Minimum wall thickness

This is required to ensure that any vessel is sufficiently rigid to withstand its

own weight, and any incidental loads. As a general guide the wall thickness of any

vessel should not less than the values given below; this includes a corrosion

allowance of 2 mm.

64

Table 2.1 Minimum wall thickness

Vessel diameter

(m)

Minimum

thickness (mm)

1 5

1 to 2 7

2 to 2.5 9

2.5 to 3.0 10

3.0 to 3.5 12

Minimum wall thickness,

tw ii

ii

Pf

DP

2

88.039.1322

40688.0

mm3538.1

Actual minimum wall thickness,

taw = tw + corrosion allowance

= 1.3538 + 2.0

= 3.3538

3.5.7 Vessel head and closure thickness

Standard torispherical heads (dished ends) are the most commonly used

end closure for vessels up to operating pressure of 15 bar.

Minimum thickness of vessel head,

t = 2.02 si

sci

CPfJ

CRP

Where Cs = stress concentration factor for torispherical heads

=

k

c

R

R3

4

1

Rc = crown radius

= Di

Rk = knuckle radius

= 0.06Rc

65

Rc = 406 mm

Rk = 24.36 mm

Cs =

36.24

4063

4

1

= 1.7706

Minimum thickness of vessel head, t =

2.07706.188.0139.1322

7706.140688.0

= 2.3767

Actual minimum wall thickness = t + corrosion allowance

= 4.3767 mm

3.5.8 Longitudinal stress

t

DP ii

h2

3539.32

40688.0h 2/2642.53 mmN

3.5.9 Circumferential stress

t

PiDiL

4

3539.34

40688.0L

2/6317.26 mmN

3.5.10 Design load

i. Dead weight of vessel

For preliminary calculations the approximate weight of a cylindrical vessel

with domed ends, and uniform wall thickness can be estimated from the following

equation:

66

3108.0 tDHgDCW mvmmvv

Where Wv = total weight of shell

Cv = 1.08 for vessels with only a few internal fittings

ρm = Density of vessel material (7750 kg/m3)

Dm = Mean diameter of vessel =

mtDi ,10 3

3103538.3406 mD

mDm 4094.0

3103538.3409.08.0581.9775008.1 vW

kNNWv 885.19672.1884

ii. Weight of tubes

gLddNW miott 22

81.9775050165.00195.0128 22 tW

kNNWt 509.161347.16509

3.5.11 Weight of insulation

Material used = 85% magnesia

Up to about 600oF (315oC), 85% magnesia has been the most popular

material. It is a mixture of magnesia and asbestos fibers so constructed that about

90% of the total volume is dead air space. Equivalents are available for situations

where asbestos is undesirable. Such insulants are applied to the equipment in the

form of slabs or blankets which are held in place with support and clips spotwelded

to the equipment. They are covered with cement to seal gaps and finished off with a

canvas that is trated for resistance to the weather. A galvanized metal outer cover

may be preferred because of its resistance to mechanical damage of the insulation.

67

Table 2.2 Insulation of 85% Magnesia or Equivalent up to 600oF

Pipe size Standard

thickness

(in)

Double

standard thickness

(in)

(in) (m)

12-33 0.3048-

0.8382

1-1/2 3

Table above was taken from Chemical Process Equipment Selection and

Design, Stanley M. Walas, page 224, table 8.22)

Insulation thickness was selected to be 1 inch (0.0254m)

Table 2.3 Thermal conductivities of insulating materials for

high temperatures

From Table 2.2 and Table 2.3, the insulation thickness and bulk density for

85% Magnesia is 1 inch and 12 lb/ft2 respectively

minchtins 0254.01

3322.19212

m

kg

ft

lb

Approximate volume of insulation

22

oinso dtdLV

68

220195.00254.00195.05 V

02570.0V m3

gVWins

kNNWins 04846.04619.4881.922.19202570.0

Total weight of Heat Exchanger:

instVT WWWW

4619.481347.165099672.1884 TW

kNNWT 4426.185638.18442

3.5.12 Pipe selection for nozzle

Pipe size for steam inlet (shell)

Material of construction = carbon steel

Density of steam inlet, ρ = 0.4872 kg/m3

Flow rate inlet, G = 0.7403 kg/s

Diameter pipe for water inlet (shell), inwaterD , = 37.053.0293 G

= 37.053.04872.07403.0293

= 325.9883 mm

Pipe size for water outlet (shell)

Material of construction = carbon steel

Density of steam outlet = 0.7045 kg/m3

Flow rate outlet, G = 0.7403 kg/s

Diameter pipe for water outlet (shell), outwaterD , = 37.053.0293 G

= 37.053.07045.07403.0293

= 284.4053 mm

Pipe size for ammonia inlet (tube)

Material of construction = stainless steel

Density of ammonia inlet = 0.8139 kg/m3

Flow rate inlet, G = 0.6842 kg/s

Diameter pipe for ammonia inlet (tube), inNHD ,3=

37.053.0293 G

69

= 37.053.08139.06842.0293

= 258.5850 mm

Pipe size for ammonia outlet (tube)

Material of construction = stainless steel

Density of ammonia outlet = 0.6098 kg/m3

Flow rate outlet, G = 0.6842 kg/s

Diameter pipe for ammonia outlet (tube), outNHD ,3=

37.053.0293 G

= 37.053.06098.06842.0293

= 287.7367 mm

3.5.13 Standard flanges

Flanges joints are used for connecting pipes and instruments to vessel, for

manholes cover and for removable vessel head when ease of access is required.

Flanged may also be used on the vessel body, when it is necessary to divide the

vessel into sections for transport or maintenance. Flanges joint are also used to

connect pipe to equipments such as pumps and valves. Flanges range in size from

a few millimeters diameter for small pipes to several meters diameter for those used

as body or head flanges on vessels.

For the design of this heat exchanger, welding-neck flanges are used. It is

because welding-neck flanges have along tapered hub between the flange ring and

the welded joint. This gradual transition of the section reduces the discontinuity

stresses between the flange and branch and increases the strength of the flange

assembly. Welding-neck flanges and branch are suitable for extreme service

conditions, where flange are likely to be subjected to temperature, shear and

vibration loads. They will normally be specified for the connections and nozzles on

process equipment. The dimensions of welding-neck flanges is chosen base on the

nominal pipe size of the nozzle pipe. All dimensions are listed below.

Standard flanges for inlet water

Diameter water inlet pipe = 325.9883 mm

Standard o.d pipe = 355.6 mm

70

Nom. size

Pipe o.d. d1

Flange Raised face

Bolting Drilling Neck

D b hi d4 f No. d2 k d3 h2 r

350 355.6 490 22 62 415 4 M20 16 22 495 438 15 12

Standard flanges for outlet water

Diameter water outlet pipe = 284.4053 mm


Nom. size

Pipe o.d. d1

Flange Raised face


D b hi d4 F No. d2 k d3 h2 r

300 323.9 440 22 62 365 4 M20 12 22 395 342 15 12

Standard flanges for inlet ammonia

Diameter ammonia inlet pipe = 258.5850 mm

Standard o.d pipe = 273 mm

Nom. size

Pipe o.d. d1

Flange Raised face


D b hi d4 f No. d2 k d3 h2 r

250 273 375 22 60 312 3 M16 12 18 335 290 15 12

Standard flanges for outlet ammonia

Diameter ammonia outlet pipe = 287.7367 mm


Nom. size

Pipe o.d. d1

Flange Raised face


D b hi d4 F No. d2 k d3 h2 r

300 323.9 440 22 62 365 4 M20 12 22 395 342 15 12

3.5.14 Design of saddles

Determination of support for a vessel will be depending on the design

temperature and pressure, vessel location and arrangement, and the internal

and external fittings. Support should be design to allow easy access to the

71

vessel for inspection and maintenance. Since heater is a horizontal

arrangement, saddle support is chosen as the support.

The saddle must be designed to withstand he load imposed by the weight of

the vessel and its contents. The design of saddle depends on the weight of

vessel, which is the weight of the heater itself. From previous calculation of

heater weight, the total weight is 18.4426 kN. From the value of weight, the

dimensions of saddle choose as referred to Figure 13.26 from Coulson &

Rochardson’s Volume 6. For outer shell diameter, Dshell is 0.406m so 0.6m is

taken since it is the smallest value and the maximum weight is not exceeded.

Vessel diamete

r (m)

Maximum weight (kN)

Dimension (m) mm

V Y C E J G t2 t1 Bolt

diameter

Bolt holes

0.6 35 0.48 0.15 0.55 0.24 0.190 0.095 6 5 20 2

3.5.15 Baffles

Baffles are used in the shell to direct the fluid flow across tube and increase

the fluid velocity. When the fluid velocity increases, it is improving the rate of heat

transfer. The assembly of baffles and tubes are hold together by support rods and

spacers. The most commonly used type of baffle is the single-segmental baffle.

Baffle cut used to specify the dimensions of a segmental baffle. Generally, baffle cut

of 20%-25% will be optimum. The value will give good heat transfer rate without

excessive drop.

Type = single segmental

Baffle diameter = 0.406 m

Nb = length of tube / inside diameter shell

= 5000 / 406

= 12.3 ≈ 13 baffles

72

Summary of design

Design pressure = 0.88 N/mm2

Design temperature = 182 oC

Material of construction = Carbon steel

Minimum thickness of cylindrical section of the shell = 3.3538 mm ≈ 4

mm

Longitudinal stress = 53.2642 N/mm2

Circumferential stress = 26.6317 N/mm2

Minimum thickness of vessel head = 4.3767 mm ≈ 5 mm

Diameter pipe for steam inlet = 325.9883 mm

Diameter pipe for steam outlet = 284.4053 mm

Diameter pipe for ammonia inlet = 258.5850 mm

Diameter pipe for ammonia outlet = 287.7367 mm

Types of baffles = Single segmental

Number of baffle segmental = 13

73

3.5 ABSORBER

3.5.1 Operating and Design Temperature and Pressure

This column operates at temperature of 66.81°C and pressure of 1 atm.

The design pressure will be 10% above the operating pressure, to avoid spurious

operation during minor process upset. The design temperature at which the design

stress is evaluated is taken as the maximum operating temperature of the material,

with due allowance for any uncertainty involved in predicting vessel wall

temperatures.

3.5.2 Materials of Construction

As one of the process material involve is ammonium nitrate, the material of

construction of the column is required to be corrosion resistant. In this case stainless

steel type 304 is selected.

Table Typical design stresses for plate

(The appropriate material standards should be consulted for particular

grades and plate thicknesses)

74

3.5.3 Column Wall Thickness

Calculating the cylindrical column wall thickness:

i

ii

Pf

DPe

2

Where e = minimum thickness required, mm

Di = internal diameter of column, mm

f = design stress, N/mm2

Pi = internal pressure, N/mm2

For corrosive process material i.e. ammonium nitrate solution, corrosion

allowance of 4 mm is included:

3.5.4 Column Head

3.5.4.1 Flat Head

Calculating the minimum thickness required:

where Cp = design constant = 0.55 for plate welded to the end of the shell

De = nominal plate diameter, mm = Di

F = design stress, N/mm2



75

3.5.4.2 Ellipsoidal head


Where J = joint factor = 1 for no joints.



3.5.4.3 Torispherical head


Where Cs = stress concentration factor =

Rc = crown radius = Di

Rk = knuckle radius = 0.06Rc

J = joint factor = 1 for no joints



76

Type of Head Minimum Thickness, e

Flat head 22mm

Ellipsoidal head 5mm

Torispherical head 5mm

By comparing the minimum thickness of these different type heads, it can

be concluded that either ellipsoidal or torispherical head are suitable to be choose

due to the economical factor since both require minimum thickness compared to flat

head.

3.5.5 The design of Column subject to Combined Loading

The main sources of load to be considered are dead weight loads and

wind. Meanwhile, the major sources of dead weight loads include vessel shell,

internal fittings (packed bed) and external fittings (ladders, platforms, piping).

3.5.5.1 Dead Weight Loads

3.5.5.1.1 Dead weight of vessel, Wv

For a steel vessel,

Where Wv = total weight of the shell, excluding internal fittings, kN

Cv = factor to account for the weight of the internal supports

= 1.15 for absorption column

Hv = height of cylindrical section, m

t = wall thickness, mm

Dm = mean diameter, m = Di + (t × 10-3)

= 1.2 + (5 × 10-3) m

= 1.205 m

77

3.5.5.1.2 Dead weight of Packed Bed, Wp

Surface area of packing, a = 95 m2/m3

Approximation volume of packed bed, Vp =

=

= 0.0283 m3

Area of packed bed, Ap = a × Vp

= 95 m2/m3× 0.0283 m3

= 2.6861 m2

For vertical column, steel platform = 1.7 kN/m2 area,

Weight of packed bed, Wp = 1.7 kN/m2 × Ap

= 1.7 kN/m2 × 2.6861 m2

= 4.5663 kN

3.5.5.1.3 Weight of External Fittings, Wfitting

External fitting used is plain steel ladder. Weight of the ladder is estimated

to be 150 N/m lengths. Therefore,

Wfitting = 150 N/m × 6 m = 900 N = 0.9 kN

Total of Dead Weight Loads = Wv + Wp + Wfittings

= (11.5804 + 4.5663 + 0.9) kN

= 17.0467 kN

3.5.5.2 Wind Loads

Wind loading will only be important on tall columns installed in the open.

Columns are usually free standing, mounted on skirt support and not attached to

structural steel work. Under this conditions, the vessel under wind loading acts as

cantilever beam.

Take wind speed, Uw = 160 km/h

To estimate the wind pressure, the following equation is used:

Pw = 0.05 Uw2

= 0.05 (160)2

= 1280 N/m2

78

Effective column diameter, Deff = Dm + 2t

= (1.2 + 0.005) m

= 1.205 m

Loading per unit length of column, Fw = Pw × Deff

= 1280 N/m2 × 1.205 m

= 1542.4 N/m

Bending moment at bottom tangent line,

= 27763.2 Nm

3.5.6 Analysis of Stresses

At bottom tangent line,

Pressure stress:

and

Where σL = longitudinal stress due to pressure, N/mm2

σh = circumferential stress due to pressure, N/mm2

P = operating pressure, N/mm2

Di = column diameter, mm

t = column wall thickness, mm

Dead weight stress (compressive):

79

Bending stress:

Where Mx = bending moment at bottom tangent line, Nmm

Iv = second moment of area of the vessel about the plane of bending

mm4

Do = outer diameter of column, mm

Di = inner diameter of column, mm

The resultant longitudinal stress:

σw is compressive and therefore negative

As no torsional shear stress, the principal stresses will be σz and σh

The radial stress is negligible ≈ (Pi/2) = 0.0507 N/mm2

The greatest difference between the principal stresses will be on the

downwind side

= σh – σz (downwind)

80

Well below the maximum allowable design stress (165 N/mm2)

3.5.7 Elastic Stability (Buckling)

The critical buckling stress,

σc

= 82.6446 N/mm2

When the vessel is not under pressure, the maximum compressive stress

will occur:

Maximum stress = σw + σb

= (0.6118 + 4.8890) N/mm2

= 5.5008 N/mm2

The maximum stress is below critical buckling stress, thus the design is

acceptable.

3.5.8 Design of Vessel Support (Skirt Design)

Type of support : Straight cylindrical skirt

θs : 90°

Material of construction : Carbon Steel

Design stress, fs : 135 N/mm2

Skirt height : 1.2 m

Young modulus : 200,000 N/mm2

Total weight of vessel : 17.0467 kN

Wind loading : 1542.4 N/m

The maximum dead weight on the skirt will occur when the vessel is full of

water.

81

Total weight:

Wtotal = Wvessel + Wapp

= (17.0467 + 66.5691) kN

= 83.6158 kN

Bending moment at skirt base:

Bending stress in skirt, σbs:

As for the first trial, take skirt thickness as the same as the thickness of the

column wall, ts = 5 mm.

Dead weight in the skirt, σws

The resulting stress in the skirt,

Maximum σs (compressive) = σbs + σws(test)

= (6.9823 + 6.9760) N/mm2

= 13.9583 N/mm2

82

Maximum σs (tensile) = σbs + σws(operating)

= (6.9823 + 1.7864) N/mm2

= 8.7687 N/mm2

General consideration for skirt design:

Take joint factor, J = 0.85

σs (tensile) < fs J sin θ

8.7687 N/mm2 < (135 N/mm2)(0.85)(sin 90°)

8.7687 N/mm2 < 114.75 N/mm2

σs (compressive) <

13.9583 N/mm2 <

13.9583 N/mm2 < 104.17 N/mm2

Both criteria are satisfied, add 2 mm for corrosion allowance,

ts = 5 mm + 2 mm = 7 mm

3.5.9 Base Ring and Anchor Bolts

Assume pitch circle diameter = 2.2 m

Circumference of bolt circle = 2200π

Recommended spacing between bolts = 600 mm

Minimum number of bolts required, Nb =

Closest multiple of 4, Nb = 12 bolts

Bending moment at base skirt, Ms =

Total weight of vessel, Wt = 17.0467 kN

Take bolt design stress, fb = 125 N/mm2

The bolt area required is given by:

83

Use bolts standard diameter = 30 mm

Use M24 bolts (BS4190:1967) root area = 353 mm2


The minimum width of the base ring:

Where Lb = base ring width, mm

fc = maximum allowable bearing pressure on the concrete

foundation

pad (typically range from 3.5 to 7 N/mm2)

Table Anchor bolt chair design

84

Actual width required:

Lb = Lr + ts +50 mm

= (76 + 7 + 50) mm

= 133 mm


Base ring thickness:

Where f’c = actual bearing pressure on base, N/mm2

fr = allowable design stress in the ring material, typically

140 N/mm2

3.5.10 Piping and Flanges Design


Where G = Fluid flowrate, kg/s

ρmix = Density of fluid mixture, kg/m3

Nozzle thickness:

Where Ps = operating pressure, N/mm2 = 0.1013 N/mm2

σ = Design stress at operating temperature, N/mm2 = 165

N/mm2

85

Pipe Flowrate, G

(kg/s)

Fluid density, ρ

(kg/m3)

Bottom inlet 2.2300 0.1584

Top inlet 0.6944 1003.6

Top outlet 1.6125 0.0356

Bottom outlet 1.3119 1276.3

Bottom inlet:

Add corrosion allowance of 4 mm,

Top inlet:


Top outlet:


86

Bottom outlet:


3.5.11 Summary of Mechanical Design

Types Packed Column

Design pressure 0.1115 N/mm2

Design Temperature 66.81°C

Cylindrical

Material Stainless Steel Type 304

Tensile strength 510 N/mm2

Design stress 165 N/mm2

Types of head Ellipsoidal @ Torispherical

Height head 0.5 m

Thickness 5 mm

Corrosion allowed 2 mm

Column weight

Dead weight 11.5804 kN

Weight of insulation NA

Weight of packed bed 4.5663 kN

Weight of external fittings 0.9 kN

Total weight 17.0467 kN

Wind loading

Loading 1542.4 N/m

Analysis stress

Dead weight stress 0.6118 N/mm2

Bending stress 4.8890 N/mm2

Critical buckling 82.6446 N/mm2

87

Vessel supports

Straight cylindrical skirt 90°

Material Carbon Steel


Skirt height 1.2 m

Total weight 83.6158 kN

Bending moment 39.979 kNm

Thickness 7 mm

Anchor bolts

Bolts 12 bolts


Area 353 m2

Bolts root diameter 30 mm

Types M24 bolts (BS4190:1967)

Piping sizing (Diameter

Optimum)

Bottom inlet 879.18 mm

Top inlet 18.79 mm

Top outlet 1290.42mm

Bottom outlet 23.93 mm

Documents

Mechanical Design [Compile Draft 1]