chapter 6 mekanik rujuk

8/3/2019 Chapter 6 Mekanik Rujuk

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CHAPTER VI

MECHANICAL DESIGN OF REACTOR

6.1 PROCESS DESCRIPTION

The reactor used in the production of phosphoric acid is continuously-stirred tank rector

(CSTR). The reaction involved is

3 H2SO4 (l) + Ca3(PO4)2 (s) + 6 H2O (l) 2 H3PO4 (l) + 3 CaSO4.2H2O(s)

Phosphoric acid and gypsum are products of the reaction. The reaction occurred at

temperature of 80C and pressure of 1.5 bar. Figure 6.1 shows the reactants and products

of the reactor.

R-1

Figure 6.1 Reactor R-100

H2SO4Ca3(PO4)2

H2O

H3PO4

CaSO4.2H2O

R-100

Stream 7 Stream 8


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6.2 MATERIAL SELECTION

The construction material for fermenter is a stainless steel type 304 which is an austenitic

iron. Type 304 has a minimum of 18% chromium, 8% nickel, and combined with a

maximum of 0.08% carbon. It is defined as a Chromium-Nickel austenitic alloy. It

contains the minimum chromium, Cr and nickel, Ni that give a stable austenitic structure.

It also has lowest price compared to other stainless steel. Type 304 has characteristics

such as good in forming and welding, corrosion or oxidation resistance, excellent

toughness, and ease of cleaning.

6.3 DESIGN SPECIFICATION

The operating pressure, PO for fermenter is 1atm (14.70 psi and 1.01325 bars)

which is same as the atmospheric pressure, Patm. Therefore, the fermenter is designed

subject to internal pressure. The operating temperature, To is 65oC (150 F). Since Po,abs

Patm , the pressure vessel is designed under internal pressure.

Nominal thickness, tnominal = 10 mm

Corrosion allowance, CA = 2 mm

Figure 6.2 shows the shape of the vessel and its dimension.

Top - ellipsoidal head with a radio of 2:1

Shellcylindrical

Bottom - ellipsoidal head with a radio of 2:1


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h/3

h/3

Di

hs

ho

ho

Figure 6.2 Dimensions of Pressure Vessel

6.3.1 Dimension calculation

From Superpro, volumetric flow rate of the outlet stream of the reactor is 546.77 L/h. At

1.5 days time, the capacity of the reactor is

V = = 19.68 m3

For chemical plant, Di: H = 1: 4, H = 4Di,

322

444

ii

ii DDD

HD

V

3368.19 iDm

Internal diameter, Di = 2.5 m = 8.2 ft = 98.43 in

HL


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Vessel height, H = 4D

= 10 m = 32.81 ft = 393.70 in

Other dimensions can be determined from the given Di and H

For the ellipsoidal height with a ratio of 2:1(major axis: minor axis), hDi 22 ,

Ellipsoidal height, f tmininD

h io 05.26251.061.244

43.98

4

Shell height, ftminininhHh os 71.2875.848.34461.24270.3932

Support line, inin

inh

hL os 89.360

3

61.24248.344

3

2

Table 6.1 Dimensions of The Vessel

Part Dimension(ft) Dimension (in) Dimension (m)

Internal diameter

Height

Ellipsoidal height

Shell height

8.20

32.81

2.05

28.71

98.43

393.70

24.61

344.48

2.50

10.00

0.63

8.75

6.4 WALL THICKNESS OF REACTOR

The wall thickness, t for each part of the pressure vessel has to be calculated in order to

get the minimum wall thickness, tmin under internal pressure.


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6.4.1 Top Ellipsoidal Head

From ASME Code Sec II Part D Table 1A, maximum allowable stress, S for stainless

steel Type 304 is 20000 psi while the joint efficiency, E is 1.0.

Operating pressure = 1.01325 bar = 14.7 psi

Design pressure, PD = Po + 0.433h

= 14.7 + 0.433(2.05)

= 15.5877 psi

From ASME Code UG-32 part (D),

Thickness, t = =

= 0.03836 in

6.4.2 Cylindrical shell


= 14.7 + 0.433(30.76)

= 28.0191 psi

From ASME Code UG-27 Part (C),

For circumferential stress, thickness, t =

=

= 0.0690 in


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For longitudinal stress, thickness, t =

=

= 0.03447 in

6.4.3 Bottom Ellipsoidal Head


= 14.7 + 0.433(32.81)

= 28.9067 psi

From ASME Code UG-32 Part (D),

Thickness, t =

=

= 0.071 in

6.4.4 Wall Thickness

tcalc = 0.071 in = 1.80mm

By considering corrosion allowance, CA of 2mm,

tuser = tcalc + CA

= 1.80mm + 2mm

= 3.80mm

tmin = tnominalCA

= 10mm2mm

= 8mm

= 0.3150 in

Table 6.2 Design Pressure and Wall Thickness of The Vessel

Part Design pressure (psi) Wall thickness (in) Wall thickness (m)

Top ellipsoidal head

Cylindrical shell

Bottom ellipsoidal head

15.5877

28.0191

28.9067

0.03836

0.06900

0.07100

0.000974

0.001753

0.001803


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6.5 MAXIMUM ALLOWABLE WORKING PRESSURE

The maximum allowable working pressure of pressure vessel, MAWP is determined by

calculating part by part under internal pressure.

6.5.1 Top ellipsoidal head


P =

=

= 127.9279 psi

MAWP vessel = 127.92790.433(2.05)

= 127.0403 psi

6.5.2 Cylindrical shell

From ASME Code UG-27 Part (C),

For circumferential stress,

P =

=

= 127.5071psi

MAWP vessel = 127.50710.433(30.76)

= 114.1880 psi

For longitudinal stress,

P =


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=

= 256.6505 psi

MAWP vessel = 256.65050.433(30.76)

= 243.3314psi

6.5.3 Bottom ellipsoidal head


P =

=

= 127.9279psi

MAWP vessel = 127.92790.433(32.81)

= 113.7212 psi

Table 6.3 MAWP of The Vessel

Part MAWP (psi)

Top ellipsoidal head

Cylindrical shell

Bottom ellipsoidal head

127.0403

114.1880

113.7212

In conclusion, MAWP vessel is 113.7212 psi, smallest value is chosen.

6.6 COMBINED LOADING

6.6.1 Primary Stresses


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For a cylindrical vessel, primary stresses which are required to achieve static equilibrium

are due to the following sources:

a) Longitudinal and circumferential stresses due to pressureb) Direct stressc) Bending stressesd) Torsional shear stresses

Some assumptions are made for the calculation of primary stress.

Assumptions :

Liquid level is three quarter of the effective length of the column. Hydrostatic pressure is considered at the bottom of the column. Liquid density in the column is equal to water (1000kg/m3) Safety factor of 10% is considered.

Column height, H = 32.81 ft = 10 m

Liquid level = (10) = 7.5m

P gage = P absP atm

Where P abs=PatmP gage = 0

Design pressure, P = gh

= (1000)(9.81)(7.5)

= 73575 N/m2

= 0.0736 N/mm2

By considering safety factor of 10%, design pressure, P = 1.10(0.0736) = 0.08096 N/mm2

(a)Longitudinal (L

) and circumferential (h

) stresses

2/0595.28

071.04

)43.98(08096.0

4mmN

t

DP iL


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2/1190.56

071.02

)43.98(08096.0

2mmN

t

DP ih

(b)Direct stress ( w )

ttDW

i

w

For a steel vessel, tDHDCW mvmvv 8.0240

With mHmtDDC vimv 4008.6,5019.20018.05001.2,15.1

NWv 6125.104438.15019.28.04008.65019.215.1240

2

/7382.08.18.11.2500

6125.10443mmNw

(c)Bending stresses (b

)

t

D

I

M i

v

b2

With M = total bending moment

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

NmM

mNDPWmNP

mtDDDmHx

WxM

effww

ioeffV

40.656492

4008.6736.3204

/736.32045037.21280,/1280

,5037.20018.025001.22,4008.6

2

2

2

2

44444 011070.05001.25037.26464

mDDI iov

2/4240.70018.0

2

5001.2

011070.0

40.65649mmNb


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(d)Torsional shear stresses ( )The torsional shear stresses can be neglected in preliminary vessel design because

these loads will normally be small. 0

Total longitudinal stresses for upwind and downwind

2/2217.364240.77382.00595.28 mmNupwind bwLz

2/3737.214240.77382.00595.28 mmNdownwind bwLz

Table 6.4 Primary Stress

Primary stress Value (N/mm2)

Longitudinal stress

Circumferential stress

Direct stress

Bending stress

28.0595

56.1190

0.7382

7.4240

6.6.2 Principal Stresses

Since 0 , 1= h = 56.1190 N/mm2

2= z = 28.0595 N/mm2

3 = 0.5P = 0.5(0.08096) = 0.04048 N/mm2

6.6.3 Maximum Allowable Stress Intensity

12 = 28.0595 N/mm2

13 = 56.0785 N/mm2

23 = 28.01902 N/mm2

The maximum allowable stress, S for stainless steel type 304 is 20000 psi or 137.895

N/mm2.

max

= 56.0785 N/mm2


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Smax

Thus, the design is safe.

6.6.4 Elastic Stability

Critical buckling stress,

p

cR

t

v

E

213

For steel at ambient temperature, E = 200000 N/mm2

with a safety factor of 12, the

244/3994.14

1.2500

8.1102102 mmN

D

t

o

c

2max

/1622.87382.04240.7 mmNwb

c max

Thus, the design is safe.

From the analysis of combined loading, the material we chose has fulfilled both

requirements of maximum stress intensity and elastic stability.

S max and c max .Therefore, the design is safe.

6.7 VESSEL SUPPORT ANALYSIS

The pressure vessel designed is a vertical cylindrical vessel. For vertical cylindrical

vessel, there are two types of skirts which are straight skirt and conical skirt. In this case,


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we chose straight skirt. A skirt support consists of a cylindrical or conical shell welded to

the base of the vessel. A flange at the bottom of the skirt transmits the load to the

foundations. Openings must be provided in the skirt for access and for any connecting

pipes where the openings are normally reinforced. The skirt may be welded to the bottom

head of the vessel, welded flush with the shell or welded to the outside of the vessel. In

this case, we chose skirt welded flush with the shell which is the most commonly used.

Skirt supports are recommended for vertical vessels as they do not impose concentrated

loads on the vessel shell and particularly suitable for use with tall columns subject to

wind loading.

6.7.1 Skirt ThicknessThe skirt thickness must be sufficient to withstand the dead-weight loads and bending

moments imposed on it by the vessel and it will not be under the vessel pressure.

The resultant stresses in the skirt are:

wsbss tensile

wsbss ecompressiv

with bending stress in the skirt,

ssss

s

bsDttD

M

4

and dead weight stress in the skirt,

sss

wsttD

W

The skirt thickness must be adequate to endure the dead-weight loads and bending

moment imposed on it by the vessel and it will not be lower the vessel pressure.

Assume skirt support height, hs = 1m, skirt thickness, mmts 20

mNW

NW

mDD

mhHx

V

is

s

/216.2345

7409.6822

8288.1

3152.813152.7

For the bending stress in the skirt,bs


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Maximum bending moment, NmWx

Ms 1085.810772

3152.8216.2345

2

22

22/698.1/311.1698047

8288.1018.0018.08288.1

1085.810774mmNmNbs

For the dead weight stress in the skirt,ws

22

/06533.0/598.65330018.0018.08288.1

7409.6822mmNmNws

Therefore, resultant stresses

2/63267.106533.0698.1 mmNtensiles

2/76333.106533.0698.1 mmNecompressivs

The skirt thickness under the worst combination of wind and dead weight loading should

follow the below design criteria:

sinJftensile ss and sin125.0

s

ss

D

tEecompressiv

Assumptions:

Weld joint factor, J = 1.0 Base angle, os 90 Maximum allowable design stress, 2/145 mmNfs Youngs modulus, 2/200000 mmNE

2/14590sin0.1145sin mmNJf

o

ss

2/06.24690sin

8288.1

018.0200000125.0sin125.0 mmN

D

tE

o

s

s

s

22 /145/63267.1sin mmNmmNwhereJftensile ss

22 /06.246/76333.1sin125.0 mmNmmNwhereD

tEecompressiv

s

ss


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Since both criteria are satisfied, the assumed value of skirt thickness mts 018.0 is

acceptable. With corrosion allowance of 4 mm, the design skirt thickness, mts 022.0 .

6.7.2 Base Ring and Anchor Belt Design

Approximate pitch circle diameter = 2.2 m.

Circumference of bolt circle = 2200

Number of bolts required, at minimum recommended bolt spacing, Nb

Closest multiple of 4 = 12 bolts

Take bolt design stress = 125 N/mm2

The bending moment at the base, Ms = 81077.1085 Nm

The weight of the vessel, W = 6822.7409 N

The area of one bolt at the root of the thread, Ab,

[ ]

[ ]


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Since the bolt root diameter is smaller than 25mm, straight skirt cannot be used and we

change it to the conical skirt.

Total compressive load on the base ring per unit length,

[ ] [ ]

Taking the bearing pressure as 5 N/mm2, the minimum width of the base ring, Lb

Take the skirt bottom diameter as 3m,

Keep the skirt thickness the same as that calculated for the cylindrical skirt. Highest

stresses will occur at the top of the skirt where the values will be close to those calculated

for the cylindrical skirt. sin 78.95 = 0.98, so this term has little effect on the designcriteria.

6.8 FLANGED JOINT DESIGN

There are several different types of flange used for various applications. In this case, the

flanged joint design we chose is lap-joint flanges. Lap-joint flanges are used in pipe work.


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They are economical when used with expensive alloy pipe like stainless steel. This is

because the flanges can be made from low-cost 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.

At stream 7, lap-joint flange is chosen.

Total mass flow rate of H2SO4 = 1.3889 kg/s, density of H2SO4 = 1840 kg/m3

For liquid flow, the optimum diameter of a flange,

From Appendix E, nominal size of flange to be used is 80mm. From the appendix,

outside diameter of pipe for 80mm flange size is 88.9mm. Thus, it can be concluded that

the outside diameter of the pipe is 88.9mm.

6.9 DESIGN DRAWING

Figure 6.3 shows the standard and detailed engineering drawing of mechanical design of

pressure vessel using Autocad software.

chapter 6 mekanik rujuk

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