Process Guidelines

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WORK IMPROVEMENT PLAN

PROCESS DESIGN GUIDELINES

WIP-SIPS-PCS-001

08/02/08 A First Issue _PCSRM _PCSGP _ENGGP

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REVISION RECORDING

Date Revision Description of Revision Prepared by Checked by Approved by

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TABLE OF CONTENTS 1. INTRODUCTION...4 2. RESPONSIBILITY ...4 3. SCOPE...4 4. PHYSICAL PROPERTIES...5 5. FORMULA ...6

6. INSULATION SPECIFICATION FOR PIPING...9

7. MATERIALS OF CONSTRUCTION ...10

8. SELECTION OF THERMODYNAMIC MODEL IN HYSYS...11

9. PIG LAUNCHER AND RECEIVER...14

10. PUMP SELECTION AND SYSTEM DESIGN ...16

11. VESSEL SELECTION AND SIZING...25

12. HEAT EXCHANGERS ...28

13. COLUMNS & TOWERS...34

14. COMPRESSORS AND VACUUM EQUIPMENT ...36

15. SAFETY SYSTEM & PSV DESIGN...38

16. PIPE SIZING ...43

17. SELECTION OF VALVES ...48

18. FLARE SYSTEM ...57

19. HEAT TRANSFER FLUID (HTF) SYSTEM DESIGN ...71

20. COOLING WATER SYSTEM DESIGN...72

21. REFRIGERATION SYSTEMS ...74

22. CHILLED WATER SYSTEM...75

23. CHILLED BRINE SYSTEM...76

24. DM WATER SYSTEM...77

25. STEAM SYSTEM...78

26. PLANT & INSTRUMENT AIR SYSTEM ...80

27. NITROGEN GENERATION SYSTEM ...82

28. INCINERATOR SYSTEM ...83

29. EFFLUENT TREATMENT SYSTEM...84

30. PROCESS CONTROLS ...86

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1. INTRODUCTION

This guideline generally outlines the methods for designing the process systems.

2. RESPONSIBILITY

The design engineer shall be responsible for carrying out the Process system design. The respective lead engineer shall counter check the correctness of calculations and design.

3. SCOPE

This guideline covers the following subjects in detail: 1) Formulas

2) Material of construction

3) Storage tank selection and sizing 4) Vessel selection and sizing 5) Pump selection and sizing 6) Line sizing and selection

7) Control valve selection and sizing 8) Compressor selection

9) Safety valve relief load calculation and selection

10) Raw water / service water / potable water system design 11) Cooling water system design

12) Chilled water / chilled brine system design 13) DM water system design

14) Steam & condensate system design 15) Service and Instrument air system design 16) Nitrogen generation / storage system design 17) Oxygen storage and system design

18) Diesel / Fuel oil / LSHS storage and handling system design 19) Flare system design and selection

20) Chemicals storage and handling system design 21) Process control

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4. PHYSICAL PROPERTIES

Property Units Water Organic _Liquids Steam Air Organic _Vapors

KJ/kg °C 4.2 1.0 - 2.5 2.0 1.0 2.0 - 4.0 Heat Capacity _{Btu/lb °F} _1.0 _{0.239 - 0.598} _0.479 _0.239 _{0.479 - 0.958} kg/m³ 1000 700 - 1500 1.29@STP Density lb/ft³ 62.29 43.6 - 94.4 0.08@STP KJ/kg 1200 - 2100 200 - 1000 Latent Heat Btu/lb 516 - 903 86 - 430 W/m °C 0.55 - 0.70 0.10 - 0.20 0.025 - 0.070 0.025 - 0.05 0.02 - 0.06 Thermal Cond. _{Btu/h ft °F} _{0.32 - 0.40} _{0.057 - 0.116} _{0.0144 - 0.040} _{0.014 - 0.029} _{0.116 - 0.35} Viscosity cP 1.8 @ 0°C **See Below 0.01 - 0.03 0.02 - 0.05 0.01 - 0.03

0.57 @ 50°C

0.28 @ 100°C

0.14 @ 200°C

Prandtl

Number 1 - 15 10-1000 1.0 0.7 0.7 - 0.8

** Viscosities of organic liquids vary widely with temperature Liquid density varies with temperature by:

3 . 0 ) (T_C T L⋅∝⋅ − ρ

Gas density can be calculated by:

T R Z P MW G ⋅ ⋅ × = ρ

Boiling Point of Water as a Function of Pressure:

(

₉

)

0.25 10 × = P T_bp ; Tbp in °C ; P in MPa Density of Metals Metal Density _(kg/m³) Aluminum 3500 Carbon Steel 7800 Galvanized Iron 6000 Stainless Steel 8000 Titanium 4000 FRP 2000

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5. FORMULA

5.1 VOLUME AND SURFACE AREA

Volume and surface area for different shapes and sections are:

Section Volume Surface area

Sphere πD³/ 6 πD²

Hemi-head πD³ / 12 πD²/ 2

S.E.head πD³ / 24 1.084D²

Ellipsoidal head πD² l / 6 2πR² + (πl2/ e) ln ((1+ e)/(1 – e))

100 – 60% F& D head 0.08467D³ 0.9286D² F & D head 2πR³ K / 3 πR² ( 1+ l2_{/ R² (2 – l/R))} Cone πD² l / 12 πDl / 2cosά Truncated cone (π l (D² + D d + d²) ) / 12 π((D + d)/2) sqrt (l²+((D – d)/2)²) 30° Truncated cone 0.227(D³ – d³) 1.57(D²– d²) Cylinder πD² l / 4 πDl . Where

l = Height of cone, depth of head or length of cylinder ά = one- half apex angle of cone

D = Large diameter of cone / diameter of head or cylinder

R = Radius

r = Knuckle radius of F & D head L = Crown radius of F & D head h = Partial depth of horizontal cylinder K,C = Coefficients

d = Small diameter of truncated cone V = Volume 2 2 1 2 1 1 R l e R r R L R L R L K − = ⎟ ⎠ ⎞ ⎜ ⎝ ⎛ ₊ ₋ ⎟ ⎠ ⎞ ⎜ ⎝ ⎛ − − =

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5.2 PARTIAL VOLUME FOR HORIZONTAL VESSELS Volume of cylinder (V1) )} cos( { R h -1 ) cos( ) cos( ) sin( 3 . 57 2 1 ∝ = ° ∝ = ° ∝ ⎟ ⎠ ⎞ ⎜ ⎝ ⎛ ∝° ₋ _∝_° _× _∝_° ⋅ ⋅ = A Degrees R L V

Volume of Head (V2) (both heads)

(

R h

)

h V₂=2×0.215× × 3 − Total Volume (V) 2 1 V V V= +

5.3 PARTIAL VOLUME FOR VERTICAL VESSELS

Volume of cylinder (V1) h D V1 2 4 π =

Volume of Head (V2) (one head)

(

R D

)

h V2=0.215× × 3 − Total Volume (V) 2 1 V V V= +

5.4 WETTED AREA FOR HORIZONTAL VESSELS

Wetted area of cylinder (A1)

(

Ar h R

)

L

R

A1= ⋅2⋅ cos(1− / ) ×

Wetted area of heads (A2) (heads are assumed to be hemispherical) (two heads)

⎟ ⎠ ⎞ ⎜ ⎝ ⎛ − − ⋅ = R H R R A2 2π 2 1

Total wetted area (A) 2 1 A A A= +

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5.5 WETTED AREA FOR VERTICAL VESSELS

Wetted area of cylinder (A1) h D A₁=π⋅ ⋅

Wetted area of heads (A2) (heads are assumed to be hemispherical) (one head) 2

2 2 R A = ⋅π⋅ Total wetted area (A)

2 1 A A A= +

Where D = Diameter of the vessel (m) R = Radius of the vessel = D/2 (m) L = Length of the vessel (m)

h = Height of liquid from vessel bottom (m) A1 = Wetted area of cylindrical portion (m²) A2 = Wetted area of heads (m²)

A = Total wetted area of the vessel (m²) V1 = Partial volume of cylindrical portion (m²) V2 = Partial volume of heads (m²)

V = Total liquid volume of the vessel (m²)

5.6 TWO PHASE DENSITY AND VISCOSITY

The density and viscosity of mixed phase fluid is found by the following method:

(

)

) 1 ( ) 1 ( / λ µ λ µ µ λ ρ λ ρ ρ λ − + = − + = + = g l h g l h g l l Q Q Q

Where Q = Volumetric flow rate (m³/h) l = Liquid

g = Gas or vapor

h = Mixed phase or homogenous phase ρ = Density (kg/m³)

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6. INSULATION SPECIFICATION FOR PIPING

6.1 INSULATION THICKNESS FOR HOT SERVICE

The maximum fluid temperature (°C) is shown as a function of the insulation thickness and pipe size.

Insulation

Thickness (mm) 0 25 40 50 65 80 90 100 115 125 140 150 165 180 190 200 Nominal Pipe

Size Maximum Fluid Temperature (°C)

1/2” 7 68 230 456 723 760 1” 7 61 205 409 653 760 1 ½” 7 48 150 302 491 707 760 2” 7 45 134 269 439 635 760 2 ½” 7 42 122 244 398 579 760 3” 7 40 111 221 361 526 710 760 4” 7 37 100 195 319 465 629 760 6” 7 35 86 164 265 386 524 676 760 8” 7 34 79 147 236 343 465 599 745 760 10” 7 33 75 136 216 312 422 544 676 760 12” 7 32 72 129 203 292 393 506 629 760 14” 7 32 70 125 196 281 379 487 605 731 760 16” 7 32 68 121 188 268 360 462 573 692 760 18” 7 31 67 117 181 258 345 442 548 661 760 20’ 7 31 66 114 176 249 333 426 527 635 750 760 Over 24” and flat surface 7 31 64 110 167 236 314 400 494 594 701 760

To obtain the correct insulation thickness, read across from the correct pipe diameter to a tabulated process temperature that is greater or equal to the actual temperatures. Read up for the insulation thickness. For example a 16” diameter pipe operating at 190°C will require 80 mm of insulation.

6.2 INSULATION MATERIAL

• Up to 343°C (650°F), magnesia is most used

• From 871°C to 1037°C (1600°F – 1900°F) a mixture of asbestos and diatomaceous earth is used

• Ceramic refractories are used for higher temperatures

• Cryogenic equipment -129°C (-200°F) employs insulation with fine pores in which air is trapped

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7. MATERIALS OF CONSTRUCTION

Material Advantage Disadvantage

Carbon Steel Low cost, easy to fabricate, abundant, most common material. Resists most alkaline environments well.

Very poor resistance to acids and stronger alkaline streams. More brittle than other

materials, especially at low temperatures.

Stainless Steel

Relatively low cost, still easy to fabricate. Resist a wider variety of environments than carbon steel. Available is many different types.

No resistance to chlorides and resistance decreases

significantly at higher temperatures. 254 SMO (Avesta)

Moderate cost, still easy to fabricate. Resistance is better over a wider range of concentrations and temperatures

compared to stainless steel.

Little resistance to chlorides, and resistance at higher temperatures could be improved.

Titanium

Very good resistance to chlorides (widely used in seawater applications). Strength allows it to be fabricated at smaller thicknesses.

While the material is moderately expensive, fabrication is difficult. Much of cost will be in welding labor. Pd stabilized

Titanium

Superior resistance to chlorides, even at higher temperatures. Is often used on sea water application where Titanium's resistance may not be acceptable.

Very expensive material and fabrication is again difficult and expensive.

Nickel Very good resistance to high temperature _{caustic streams.} Moderate to high expense. _{Difficult to weld.}

Hastelloy Alloy Very wide range to choose from. Some have been specifically developed for acid services where other materials have failed.

Fairly expensive alloys. Their use must be justified. Most are easy to weld.

Graphite One of the few materials capable of withstanding weak HCl streams.

Brittle, very expensive, and very difficult to fabricate. Some stream components have been known to diffusion through some types of graphites. Tantalum Superior resistance to very harsh services _{where no other material is acceptable.} Extremely expensive, must be _{absolutely necessary.}

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8. SELECTION OF THERMODYNAMIC MODEL IN HYSYS

Application Type Model Applicable Range Remarks

OIL & GAS

Peng-Robinson

(PENG-ROB) Pressures All Temperatures All

Should not be used if polar components such as Alcohols are present.

Reservoir Systems

Equations of state for high pressure hydrocarbon

applications Redlich-Kwong-_Soave (RK-SOAVE)

All

Pressures Temperatures All

Should not be used if polar components such as Alcohols are present.

Peng-Robinson

Should not be used if polar components such as Alcohols are present

Platform Separation

All

Peng-Robinson

Transportation of oil and gas by pipeline

All

Should not be used if polar components such as Alcohols are present REFINERY Braun K-10 (BK10) Suited for vacuum and low pressure applications Applicable range is 133 -800o_{K. Can be} used up to 1100o_K

Gives fast and acceptable answers. Should be use as a first attempt only. This model should not be used for systems containing Hydrogen. Chao-Seader

(CHAO-SEA) < 140 atm 200-533o_K

This model should not be used for systems containing Hydrogen. If light ends dominate, Equation of State models can be used. Low Pressure

applications (up to several atm) eg: Vacuum tower & crude tower

Petroleum Correlation Models

Grayson < 210 atm 200-700o_K

This model should be used for systems containing Hydrogen. If light ends dominate, Equation of State models can be used. Chao-Seader

(CHAO-SEA) < 140 atm 200-533o_K

This model should not be used for systems containing Hydrogen. If light ends dominate, Equation of State models can be used. Petroleum

Correlation Models

Grayson < 210 atm 200-700o_K

This model should be used for systems containing Hydrogen. If light ends dominate, Equation of State models can be used. Peng-Robinson

Medium Pressure applications (up to several tens of atm) Coker Main Fractionator FCC Main Fractionator Equation Of State Models Redlich-Kwong-Soave (RK-SOAVE) All

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Petroleum

Correlation Models Grayson < 210 atm 200-700oK

This model should be used for systems containing Hydrogen. If light ends dominate, Equation of State models can be used Peng-Robinson

Should not be used if polar components such as Alcohols are present Hydrogen Rich Applications eg: Reformer, Hydrofiner, Hydrotreaters &

Hydro-Desulfurisers Equation Of State _Models

Redlich-Kwong-Soave

(RK-SOAVE)

All

Peng-Robinson

Lube Oil Unit, De-Asphalting Unit Equation Of State Models Redlich-Kwong-Soave (RK-SOAVE) All

GAS PROCESSING

Peng-Robinson

Should not be used if polar components such as Alcohols are present Hydrocarbon separations De-methanizer C3-splitter Equation Of State Models Redlich-Kwong-Soave (RK-SOAVE) All

Peng-Robinson

Cryogenic gas

processing Equation Of State Models Redlich-Kwong-Soave

(RK-SOAVE)

All

Should not be used if polar components such as Alcohols are present Peng Robinson MHV2 equation of state (PRMHV2) Accurate up

to 150 bar Temperatures All

Can be used for mixture of polar and non polar components Peng Robinson Wong-Sandler Equation Of State (PRWS) Accurate up

Can be used for mixture of polar and non polar components Predictive

Soave-Redlich-Kwong (PSRK)

All

Can be used for mixture of polar and non polar components Redlich Kwong MHV2 equation of state (RKSMHV2) Accurate up

Can be used for mixture of polar and non polar components Gas Dehydration with glycols Flexible and Predictive Equation of State Model Redlich Kwong Wong-Sandler Equation Of State (RKSWS) Accurate up

Can be used for mixture of polar and non polar components

PETROCHEMICALS

Chao-Seader

(CHAO-SEA) < 140 atm 200-533o_K This model should not be _{used for systems containing}

Hydrogen. Ethylene Plant Main Fractionator Petroleum Correlation Models

Grayson < 210 atm 200-700o_K This model should be used _{for systems containing}

Hydrogen. Ethylene Plant

Light Hydrocarbon

Equation Of State

Models Peng-Robinson (PENG-ROB) Pressures All Temperatures All

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Separation Train Redlich-Kwong-Soave

(RK-SOAVE)

All

Peng-Robinson

Ethylene Plant

Quench Tower Equation Of State Models Redlich-Kwong-Soave

(RK-SOAVE)

All

Should not be used if polar components such as Alcohols are present NRTL Low pressures up to 10 atm. No component should be close to its critical temperature. UNIFAC Low pressures up to 10 atm. No component should be close to its critical temperature. Aromatics (eg:

BTX Extraction) Liquid Activity Coefficients

UNIQUAC Low pressures up to 10 atm. No component should be close to its critical temperature. Peng-Robinson

Should not be used if polar components such as Alcohols are present Substituted hydrocarbons, VCM Plant & Acrylo Nitrile Plant Equation Of State Models Redlich-Kwong-Soave (RK-SOAVE) All

Should not be used if polar components such as Alcohols are present NRTL Low pressures up to 10 atm. No component should be close to its critical temperature. UNIFAC Low pressures up to 10 atm. No component should be close to its critical temperature. Ether Production eg: MTBE, ETBE, TAME Liquid Activity Coefficients(very sensitive to parameters) UNIQUAC Low pressures up to 10 atm. No component should be close to its critical temperature. Peng-Robinson

Should not be used if polar components such as Alcohols are present Equation Of State Models Redlich-Kwong-Soave (RK-SOAVE) All

Should not be used if polar components such as Alcohols are present NRTL Low pressures up to 10 atm. No component should be close to its critical temperature. UNIFAC Low pressures up to 10 atm. No component should be close to its critical temperature. Ethyl Benzene and Styrene Plants Liquid Activity Coefficients UNIQUAC Low pressures up to 10 atm. No component should be close to its critical temperature.

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9. PIG LAUNCHER AND RECEIVER

Typical diameters of Major Barrel and Pipework:

Pipeline

diameter Bypass line Kicker line Balance line Drain line Major barrel

4” 3” 2” 2” 2” 6” 6” 4” 2” 2” 2” 8” 8” 4” – 6” 4” 2” 2” 10” 10” 6” 4” 2” 2” 12” 12” 6” – 8” 4” 2” 2” 16” 14” 6” – 10” 4” 2” 2” 16” 16” 8” – 12” 6” 4” 4” 18” 18” 10” – 12” 8” 4” 4” 20” 20” 10” – 16” 8” 4” 4” 24” 24” 12” – 18” 8” 4” 4” 28” 28” 16” – 20” 10” 4” 4” 32” 30” 16” – 24” 10” 4” 4” 36” 32” 16” – 24” 10” 4” 4” 36” 36” 18” – 28” 12” 4” 4” 40” 38” 20” – 28” 12” 4” 4” 42” 40” 20” – 32” 12” 4” 4” 44” 42” 20” – 36” 16” 4” 4” 46” 48” 24” – 36” 18” 4” 4” 52” 56” 32” – 40” 20” 4” 4” 60” Kiker line B_L A_L LAUNCHER Kiker line B_R A_R RECEIVER

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Barrel lengths for Intelligent Pigs:

Approx. minimum barrel length (m)

(Notes 1 & 2) Launcher Receiver Pipeline diameter Approx. maximum tool length (m) (Note 2) Approx. maximum tool weight (kg) (Note 3) AL BL AR BR 4” 2.8 60 2.8 0.5 2.8 2.8 6” 2.8 90 2.8 1.5 2.8 2.8 8” 3.9 170 4.1 1.5 3.9 3.9 10” 4.3 300 4.3 1.5 4.3 4.3 12” 4.3 365 4.3 1.5 4.3 4.3 14” 4.8 380 4.8 1.5 4.8 4.8 16” 5.1 700 5.1 1.5 5.1 5.1 18” 5.1 810 5.1 1.5 5.1 5.1 20” 5.1 840 5.1 1.5 5.1 5.1 24” 5.7 1600 5.7 1.5 5.7 5.7 28” 5.8 2000 5.8 1.5 5.8 5.8 30” 6.0 2000 6.0 1.5 6.0 6.0 32” 6.6 2270 6.6 1.5 6.6 6.6 36” 6.6 3560 6.6 1.5 5.3 6.6 38” 6.6 3600 6.6 1.5 5.5 6.6 40” 6.6 4090 6.6 1.5 5.5 6.6 42” 6.6 4550 6.6 1.5 6.4 6.6 48” 6.6 Note 4 6.6 1.5 6.6 6.6 56” 6.6 Note 4 6.6 1.5 6.6 6.6 NOTES:

1) Refer the figure above for details

2) The lengths are extreme figures. To be checked with supplier for accurate dimensions 3) The weight is indicative for the pig only, excluding the weight of lifting/loading trolley 4) To be checked with supplier

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10. PUMP SELECTION AND SYSTEM DESIGN

10.1 PUMP SELECTION

Service Type of Pump

Very high head at low flow Reciprocating

Metering small flows (< 1 m³/h) Reciprocating (plunger or diaphragm) Viscous fluid (> 1000 cP) Rotary gear or screw

Non-Newtonian fluid Screw

Suspension of crystals or fragile solids Rotary lobe

Entrained gas (> 2 vol%) Rotary or diaphragm • Centrifugal pumps:

Single stage for 3.4 - 1134 m³/h (15 - 5000 GPM) & 152 m maximum head Multi Stage for 4.6 - 2500 m³/h (20 - 11,000 GPM) & 1675 m maximum head

Efficiencies of 45% at 23 m³/h (100 GPM), 70% at 113 m³/h (500 GPM), 80% at 2270 m³/h (10,000 GPM).

• Axial pumps can be used for flows of 4.6 - 22680 m³/h (20 - 100,000 GPM) Expect heads up to 12 m and efficiencies of about 65-85%

• Rotary pumps can be used for flows of 0.23 - 1134 m³/h (1 - 5000 GPM) Expect heads up to 15,200 m (50,000 ft) and efficiencies of about 50 - 80% • Reciprocating pumps can be used for 2.3 - 22680 m³/h (10 - 100,000 GPM)

Expect heads up to 300,000 m (1,000,000 ft).

Efficiencies:70% at 7.46 kW (10 hp), 85% at 37.3 kW (50 hp) and 90% at 373 kW (500 hp)

10.2 PUMP PERFORMANCE WITH IMPELLER AND SPEED CHANGE

Diameter change 3 1 2 1 2 2 1 2 1 2 1 2 1 2 ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ × = ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ × = × = D D BHP BHP D D H H D D Q Q Speed change 3 1 2 1 2 2 1 2 1 2 1 2 1 2 ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ × = ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ × = × = N N BHP BHP N N H H N N Q Q

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Dia & Speed change 3 1 2 1 2 1 2 2 1 2 1 2 1 2 1 2 1 2 1 2 ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ × × = ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ × × = ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ × × = N N D D BHP BHP N N D D H H N N D D Q Q

10.3 EFFECT OF VISCOSITY ON PUMP PERFORMANCE

• Alternate for centrifugal pump (rotary pump) to be considered when fluid viscosity is above 220 cSt (or 2500 SSU). Small pumps become impractical above 220 cSt.

• Generally centrifugal pumps are limited to about 4000 SSU max. viscosity

• Correction factors to head, capacity and efficiency should be applied when viscosity is above 70 SSU.

• Viscosity correction calculation:

Viscous capacity (gpm) Qvis = Qwater* Cq Cq=0.95-0.9 @200-1000 cSt Viscous head (ft) Hvis = Hwater* Ch Ch=0.85-0.95 @200-1000 cSt Viscous efficiency Evis = Ewater* Ce Ce=0.55-0.75 @200-1000 cSt Correction factors (Cq,Ch,Ce)

BHP @normal service = (Q*H*SG)/(2.31*1750*E)

BHP @viscous service = (Qvis*Hvis*SG)/(2.31*1750*Evis)

10.4 EFFECT OF VAPOR OR GASES

1) Most centrifugal pumps can handle up to 3 vol. of vapor per 100 vol. of liquid with the max.

limit at 7 - 8%.

2) If vapor is suspected in the pump suction area, specify any of the following: - Self-priming type pump

- Increase suction pressure - Specify low speed pump

- Specify that pump casing should be capable of accepting an oversize impeller to counter the resulting loss in head and capacity.

10.5 MINIMUM FLOW BYPASS

1) Single stage pumps, bypass = 15 - 25 % Multistage pumps, bypass = 25 - 35 % Worthington suggests 30 gpm per 100 HP

2) The temperature rise in a turbine regenerative pump is much greater than in a conventional centrifugal pump, since the HP input increases as the flow rate thro the turbine pump is decreased, where as the HP input to a centrifugal pump decreases as flow is reduced. Consequently a turbine pump is always provided with bypass and also a relief valve

3) The increase in liquid temperature at low flow rates may be calculated by assuming that all the HP shown on the pump curve at the desired capacity is being converted to heat, except that which is used to deliver the small capacity (by the following formulae).

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(

)

Cp Sg Q Sg H Q BHP T fromcurve × × × × × × × − = ∆ 33 . 8 4 . 42 000252 . 0 At shut-off

(

)

Cp W BHP T atshutoff × × = ∆ ⋅ 42.4

Where ∆T = Temp diff (°F), Q = gpm, H = head (ft), Sg = Specific gravity; Cp = Specific heat capacity (Btu/lb°F), W = weight of fluid in pump (lb) 4) Minimum flow orifice sizing can be done by the following formulae:

⎟ ⎠ ⎞ ⎜ ⎝ ⎛∆ × × × = Sg P d K Q 29.8 2

Where Q = gpm, K = orifice disc coef =0.65, d = orifice dia (inch), ∆P = pres. drop (psi), Sg = Specific Gravity

10.6 MINIMUM NPSH® FOR CENTRIFUGAL PUMPS

Pump capacity (m³/h) NPSHr (m) Pump capacity (m³/h) NPSHr (m)

Up to 11.4 1.5 159 - 227 3.7 – 4.3 11.4 – 22.7 1.8 227 - 455 4.3 – 6.1 22.7 – 45.5 2.1 – 2.4 454 - 568 6.1 – 7.3 45.5 - 91 2.7 – 3.0 > 568 7.6 91 - 159 3.0 – 3.7

Required NPSH can be reduced as follows: - Use a double suction pump - Use a slower speed

- Use smaller pumps in parallel - Use an oversize pump - Use impeller with inducer

10.7 PUMP CENTERLINE ELEVATION FOR CENTRIFUGAL PUMPS

Pump capacity (m³/h)

Centerline

Elevation (m) Pump capacity (m³/h)

Centerline Elevation (m)

Up to 45.5 0.76 227 - 2270 1.1

45.5 – 227 0.91 2270 - 4545 1.4

10.8 CENTRIFUGAL PUMP EFFICIENCY

Pump capacity (m³/h) Efficiency (%) Pump capacity (m³/h) Efficiency (%)

Up to 11.4 20 – 40% 45.5 – 114 50 – 75%

11.4 – 22.7 30 – 50% 114 - 227 60 – 80%

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10.9 CENTRIFUGAL PUMP MOTOR RATING

Shaft Power (BkW) Motor Rating

Up to 22.4 1.25 x BkW 22.4 – 74.6 1.15 x BkW > 74.6 1.10 x BkW

10.10 STEAM TURBINE DRIVEN CENTRIFUGALPUMPS

Shaft Power (BkW) Efficiency (%) Shaft Power (BkW) Efficiency (%)

37.3 33 373 58

74.6 37 746 64

149.2 45

10.11 PRESSURE DROP FOR EQUIPMENT AND PIPIING

Equipment / Piping Item Pressure Drop (bar)

Exchangers / Air coolers / Double pipes 0.7

Pump suction screen 0.07

Rotary & turbine flow meters 0.5

Flow orifice 0.2

Pressure drop Pressure drop

Pump suction 0.23 - 0.68 bar/km

Pump discharge

Carbon Steel Pipe Alloy Steel Pipe

0 - 57 m³/h 5.7 – 22.6 bar/km 13.6 – 33.9 bar/km

57 - 159 m³/h 3.4 – 15.8 bar/km 13.6 – 33.9 bar/km

> 159 m³/h 2 – 9.1 bar/km 4.5 – 15.8 bar/km

Higher velocity is considered for bigger pipes and for higher operating pressures.

10.12 MULTIPLYING FACTOR FOR EQUIVALENT LENGTH

Pipe dia Approx. straight length (m)

30 60 150

up to 3" 1.9 1.6 1.2

4" 2.2 1.8 1.3 6" 2.7 2.1 1.4

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10.13 PUMP SEALS

In order to prevent leakage of pump fluid through the clearance between the pump shaft and casing, two types of sealing device are generally used in pumps namely; (1) Gland Packing; (2) Mechanical Seal

Gland Packing:

This is a simple packing by a sealing device which is a rope like material wound between the pump shaft and casing. Frequent changeover of packing is required to prevent leakage. This is used for domestic and water applications where leakage is not a concern.

Mechanical Seals

Single mechanical seals are generally used to arrest leakage, without frequent maintenance. Double mechanical seals are sometimes preferred in toxic, flammable and corrosive services. 1) Packed seal (Mechanical seal)

Packed seal does not require an external flushing liquid if the pumped liquid has lubricating properties at the seal conditions. An external flushing system is required for the following:

- For vacuum service to prevent air intake. - For abrasive containing fluids.

- For handling volatile fluids which vaporize at operating conditions and present a fire hazard.

- For pumping toxic or corrosive liquids. 2) Seal piping

- Seal cooler duty can be estimated by for a seal fluid flow of 0.5 m³/h per seal. - For dead end seals the seal fluid flow can be 0.23 m³/h per seal

- Seal fluid pressure can be 0.7 – 1.72 bar above packing box pressure. Turbine or pump seal, gland, stuffing box cooling water rate shall be 1-2 m³/h. Refer API-610 to understand more about the type of flushing plans for each service.

10.14 CONTROL VALVE PRESSURE DROP FOR CENTRIFUGAL PUMPS

1) Minimum pressure drop for a control valve is 0.7 bar @design flow rate.

2) Generally CV loss shall be taken as 25% of the total system friction loss or 10% of operating pressure

3) More detailed approach for CV pressure drop is:

s f

v A dP B dP

dP = × + ×

Where: dPv = Control valve pressure drop dPf = Frictional pressure drop of system

dPs = Pressure differential due to static head of system A, B = Multiplication factors (as given below)

Qd = Design / rated flow rate Qm = Maximum process flow rate

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Pump Over design Factor (Qd / Qm) A B 0 - 5 % 0.294 0.059 10% 0.468 0.094 15% 0.658 0.132 20% 0.875 0.175 25% 0.950 0.200

10.15 PUMP SYSTEM CURVE

Where:

dPv = Control valve pressure drop at maximum flow Qm

dPo = Control valve pressure drop at full open condition at design flow Qd

10.16 PUMP CALCULATION Hso Hd dPv dPf dPs dPo Qmin Qn Qm Qd Pa dPf1 H1 H2 dPf2 Pb PS Pd H3

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Where: Pa = Suction vessel minimum operating pressure (barg) Pb = Discharge vessel maximum operating pressure (barg)

Pv = Vapor pressure of fluid at maximum operating temperature (bara) Ps = Pump suction pressure (barg)

Pd = Pump discharge pressure (barg)

Sg = Specific Gravity of the fluid at operating conditions dPf1 = Suction piping frictional loss (bar)

dPf2 = Discharge piping frictional loss (bar)

H1 = Suction vessel minimum liquid level (LSLL) elevation (m) (bar) H2 = Pump suction nozzle elevation (m) (bar)

H3 = Discharge vessel maximum liquid level (LSHH) elevation (m) (bar) To convert elevation in meters to pressure unit = H / 10.2 * Sg (in bar)

Suction pressure (Ps) = Pa + H1 – H2 – dPf1 (barg) Discharge pressure (Pd) = Pb + H3 – H2 + dPf2 (barg) Pump differential pressure (DP) = Pd – Ps (bar)

Pump head (h) = DP *10.2 / Sg (m)

Available NPSH (NPSHa) = (Pa + 14.7) – Pv – dPf1 + (H1 - H2) (bara)

= NPSHa *10.2 / Sg (m)

General considerations:

1) RVP is lower than the True vapor pressure (TVP or Absolute vapor pressure) of hydrocarbon liquids. Correction chart is used to find out TVP from RVP at operating temperature. TVP is used for pump calculations

2) Calculated NPSHa - 1 m (less 1 m) should be given in the process datasheet. The difference between vendor NPSHr and datasheet NPSHa should not be less than 1 m. 3) In absence of data pump suction elevation can be considered as 0.6 m for smaller pumps

and 1.0 m for larger pumps (> 200 m³/h).

4) In absence of data take 0.1 bar line loss for suction piping or 0.05 bar for smaller pumps. 5) Spare single stage pumps in service over 260°C and spare multistage pumps in service

over 200°C must be kept warm for quick start-up, by providing a bypass around check valve. Needle valve to be provided for the bypass.

6) Pumping capacity over design: Column overhead pumps 20%

Product and transfer pumps 10%

Design pressure of pump = Design pressure of suction vessel (PSV set) + Max. suction static

head + Max. pump differential pressure Pump Power calculation

Power estimates for pumping liquids

BkW = { Flow (m³/h) x Head (m) x Sp.Gr. } / {367 x Efficiency} BHP = { Flow (US gpm) x Head (ft) x Sp. Gr. } / {3960 x Efficiency}

**Efficiency expressed as a fraction in these relations

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Efficiency = 80-0.2855F+0.000378FG-0.000000238FG^2+0.000539F^2-0.000000639(F^2)G + 0.0000000004(F^2)(G^2) Where; Efficiency is in fraction form, F = developed head (ft), G = flow (gpm)

Ranges of applicability are F = 15 - 91 m and G = 23 – 230 m³/h

10.17 PUMP OVERDESIGN FACTOR

Service Over design Factor Service Over design Factor

Product 10% Sour water drain 25%

Reflux 15% Intermittent 0%

Circulating reflux 5% Utilities 10%

10.18 RECIPROCATING PUMP

a) Simplified formulas for sizing dampers (for pump speeds up to 100 rpm)

Pump type Damper volume, US Gallons

Simplex, single-acting V = 5 x discharge rate in gpm / rpm Simplex, double-acting V = 2.5 x discharge rate in gpm / rpm

Duplex, double-acting V = 1.3 x discharge rate in gpm / rpm Triplex, single & double-acting V = 0.45 x discharge rate in gpm / rpm For pump speeds above 100 rpm multiply the above volumes by (Pump rpm / 100).

b) Calculation of line shock pressure due to valve closure

The magnitude of line shock can be calculated by the following formula:

g v V V Sg P P × − × × = 144 ) (

Where,P = Increase in pressure due to shock (psi)

Vp = Velocity of pressure wave propagation in pipe (approx 4000 ft/s in small pipes)

Sg = Specific gravity of liquid (lb/ft³)

V = Velocity of liquid in pipe before valve closure (ft/s)

v = Velocity of liquid in pipe at an interval equivalent to the time that a pressure wave will travel up the pipe and back after the valve starts closing (ft/s)

G = Acceleration of gravity, 32 (ft/s²) Example:-

A 3" line 1000 ft long is carrying water at a pressure of 250 psi and at a velocity of 10 ft/s is suddenly shut-off by a valve closing in 0.3 sec. Assuming that the velocity of the pressure wave in the pipe is approx 4000 ft/s, it would take 0.5 sec for the wave to travel up the pipe and back. This is slower than the time required for the valve to close and consequently v (the velocity thro the valve after the pressure wave has made one complete cycle) is equal to zero.

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P = 62.5 * 4000 * (10 - 0) / (144 * 32.2) = 540 psi

Total surge pressure = Original pressure in pipe + shock pressure = 250 + 540 = 790 psi

In general water hammer will occur when the total surge pressure exceeds twice the static pressure. Therefore, in the above example, water hammer will occur.

c) Calculation of damper volume required to reduce shock

Dampers to absorb the shock of fast closing valves are mounted just upstream of the valve. To calculate the minimum surge volume, the following equation may be used.

(

)

(

)

⎥_⎦⎤ ⎢ ⎣ ⎡ − − × × × = 1 2 2 (0.005 ) 004 . 0 P P T L P R A

Where, A = Surge volume required (US gpm) R = Pipe flow rate (US gpm)

T = Normal closing time of valve in sec. (T=0 for instantaneous closure valves)

L = Length of pipe (ft)

P1 = Actual flow pressure at the inlet of valve (source pres. - line DP) (psi) P2 = Upper pressure limit which the surge should be limited to in absorbing

the decelerating flow on valve closure. This should be set at 1.5 times the static pressure in the line when the valve is closed and the liquid is at rest (psi)

Example:

In a 3"-40 sch line 1000 ft long water @ 230 gpm at 10 ft/s is flowing at 250 psi. Fric. Drop is 4.5 psi/100ft. Valve closure time is 0.3 sec.

P1 = 250 - (4.5*1000/100) = 205 psi

P2 = 250 * 1.5 = 375 psi

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11. VESSEL SELECTION AND SIZING

11.1 STORAGE TANKS CAPACITY

1) For fixed and roofless tanks the working capacity should be the volume between the top of the suction nozzle and the maximum safe working level in the tank.

2) For floating roof tanks the working capacity should be the volume between the maximum highest safe position for the roof and the minimum allowed position for the roof.

3) For study purposes the working capacity will be multiplied by 1.05 for fixed roof tanks and 1.10 for floating roof tanks to obtain nominal tank capacities for cost estimating purposes.

11.2 TYPE OF ROOF

U.S. environmental protection agency requires the following for storage of hydrocarbon in petroleum refineries:

1) Hydrocarbons with a natural vapor pressure of 1.5 psia (0.103 bara) or less at storage temperature may be stored in a freely vented fixed roof tank

2) Hydrocarbons with a natural vapor pressure between 1.5 psia to 11.0 psia (0.103 bara to 0.76 bara) at storage temperature may be stored in floating roof tanks

3) Hydrocarbons with a natural vapor pressure in excess of 11.0 psia (0.76 bara) at storage temperature should be stored in a fixed roof tank with a vapor recovery or refrigeration system.

4) The flammability of the stored material, regulations governing the emission of the vapor to the atmosphere together with cost of the product lost due to evaporation, should be considered.

11.3 TYPE OF STORAGE VESSELS

1) For less than 3.8 m³ (1000 gallons) vertical tanks on legs can be used.

2) Between 3.8 m³ and 38 m³ (1000 to 10,000 gallons) horizontal tanks on concrete supports can be used.

3) Beyond 38 m³ (10,000 gallons) vertical tanks on concrete pads can be used. 4) Liquids with low vapor pressures, tanks with floating roofs can be used. 5) Raw material feed tanks are often specified for 30 days feed supplies

6) Storage tank capacity should be at 1.5 times the capacity of mobile supply vessels. 7) For example, 28.4 m³ (7500 gallon) tanker truck, 130 m³ (34,500 gallon) rail cars 8) Liquid drums are usually horizontal. Gas/Liquid separators are usually vertical 9) Optimum Length/Diameter ratio is usually 3, range is 2.5 to 5

11.4 HOLDUP TIME

1) Holdup time is 5 minutes for half full reflux drums and gas/liquid separators 2) Design for a 5-10 minute holdup for drums feeding another column 3) For drums feeding a furnace, a holdup of 30 minutes is a good estimate

4) Knockout drum in front of compressors should be designed for a holdup of 10 times the liquid volume passing per minute.

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11.5 VELOCITY CRITERIA

1) Liquid/Liquid separators should be designed for settling velocities of 2-3 inch/min (0.00085 m/s to 0.0013 m/s)

2) Gas velocities in gas/liquid separators

(

−1

)

⋅ = V L k Velocity ρ ρ

Where k = 0.35 with horizontal mesh deentrainers 0.167 with vertical mesh deentrainers

0.1 without mesh deentrainers

Velocity is in ft/s

A 6” (150 mm) mesh pad thickness is very popular for such vessels

3) For positive pressure separations, disengagement spaces of 6”-18” (150 mm to 450 mm) before the mesh pad and 12 inches (300 mm) after the pad are generally suitable

11.6 DESIGN CONDITIONS

1) The design pressure for atmospheric storage tanks is +100 mmWC (or full of water) / -50 mmWC

2) Design pressure is 10% or 0.69 bar to 1.7 bar above the maximum operating pressure, whichever is greater, as applicable

3) The maximum operating pressure is taken as 1.7 bar (25 psi) above the normal operation pressure, as applicable

4) For vacuum operations, design pressures are 1 barg (15 psig) to full vacuum

5) For systems with maximum operating temperature (MOT) between -30°C and 345°C, design temperatures is typically MOT + 15°C (to 25°C as applicable). Above this range the margin increases

6) Minimum design temperature is 0°C to 5°C less than the minimum operating temperature 7) Minimum wall thicknesses for maintaining tank structure are:

6.4 mm (0.25 in) for 1.07 m (42 in) diameter and under 8.1 mm (0.32 in) for 1.07 m - 1.52 m (42-60 in) diameter 9.7 mm (0.38 in) for diameters over 1.52 m (60 in)

8) Allowable working stresses are taken as 1/4 of the ultimate strength of the material

11.7 MECHANICAL DESIGN

1) Thickness calculation based on pressure and radius is given by:

(Pressure) * (outer Radius) Thickness =

(Allowable stress) * (Weld Efficiency) – 0.6 (Pressure)

Where : pressure = psig, radius = inch, stress = psi, corrosion allowance = inch ** Weld efficiency can usually be taken as 0.85 for initial design work

2) Guidelines for corrosion allowances are as follows: 0.35 in (9 mm) for known corrosive fluids, 0.15 in (4 mm) for non-corrosive fluids, and 0.06 in (1.5 mm) for steam drums and air

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receivers

3) Generally 1.5 mm to 3.0 mm is considered as the minimum corrosion allowance (CA) for Carbon Steel equipment/piping while Nil corrosion allowance is considered for Stainless Steel and non-metallic equipment/piping

11.8 SIZING EXPLOSION HATCHES

e W W =3600× '

Where W = Required relieving capacity (lb/h) W' = Weight of air and gas in the vessel (lb) e = Time to attain maximum pressure (sec)

Time to attain maximum pressure for mixture of gases and air at 1 atm & 150°F is 0.01 sec for H2, 0.045 sec for ethane, 0.056 for propane, 0.06 for hexane and naphtha, 0.0117 for acetylene, 0.06 for benzene, 0.10 for toluene.

The required open area, A, to discharge the relieving capacity, W, using a discharge coefficient of 0.8 is: 1 245 / P M T Z W A × ⋅ × =

Where A = Discharge area (in²) P1 = Relieving pressure (psia) T = Initial temperature (R)

Z = Compressibility factor, at P1 & T M = Mixture MW

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12. HEAT EXCHANGERS

Some of the important design tips for heat exchanger design are:

1) Fluids that are corrosive, fouling, scaling, high pressure drop or under high pressure are usually placed in tube side

2) Hot, viscous and condensing fluids are typically placed on the shell side

3) Pressure drops are about 1.5 psi (0.1 bar) for boiling/vaporization and 3-10 psi (0.2-0.7 bar) for other services

4) The minimum approach temperature for shell and tube exchangers is about 20°F (10 °C) for fluids and 10°F (5°C) for refrigerants.

5) Cooling tower water is typically available at a maximum temperature of 90°F (30°C) and should be returned to the tower no higher than 115°F (45°C)

6) Double pipe heat exchangers may be a good choice for areas from 100 to 200 ft2 (9.3-18.6 m2)

7) Spiral heat exchangers are often used to slurry interchangers and other services containing solids

8) Plate heat exchanger with gaskets can be used up to 320°F (160°C) and are often used for interchanging duties due to their high efficiencies and ability to "cross" temperatures. 9) For the heat exchanger equation, Q = UAF (LMTD), use F = 0.9 when charts for the LMTD

correction factor are not available

10) Shell and Tube heat transfer coefficient for estimation purposes can be found in many reference books

11) Most commonly used tubes are ¾” (19 mm) outer diameter on a 1” triangular spacing at 16 ft (4.9 m) long

12) A 1 ft (300 mm) shell will contain about 100 ft2 (9.3 m2) A 2 ft (600 mm) shell will contain about 400 ft2 (37.2 m2) A 3 ft (900 mm) shell will contain about 1100 ft2 (102 m2)

13) Typical velocities in the tubes should be 3-10 ft/s (1-3 m/s) for liquids and 30-100 ft/s (9-30 m/s) for gases

12.1 DESIGN MARGINE

The following design margins shall be adopted as applicable:

1) Heat exchangers shall be designed with 10% margin on area while using the normal heat duty.

2) While designing heat exchangers, 10% margin on normal duty shall be considered and the exchanger shall be designed with minimum over design factor (say 2% to 5%) on area. 3) If the design duty (10% to 15% over design factor) is used for sizing the heat exchangers,

then the exchanger shall be designed with minimum over design factor (say 2% to 5%) on area.

12.2 EVAPORATORS

1) When the boiling point rise is appreciable, the economic number of effects in series with forward feed is 4-6.

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12.3 HEAT EXCHANGER DETAILS

1) TEMA shell types:

2) Baffle cut orientation; parallel, perpendicular and 45° with reference to shell nozzle

3) Tube length : Length up to the tangent point of the outermost tube for U-tubes. The length in the tube sheets should be included. Effective area is excluding the tube sheet.

4) Shell inlet nozzle location

5) U-tube nozzle location

Parallel

0° Perpendicular_90° 45°

U-tube exchanger

Overall tube length Total tube length

code 0 Vertical Horizontal code 2 code 1

Code 0, normal Code 1, in front of Code 2, behind U

ACU

TEMA X TEMA G TEMA H

TEMA J21 TEMA J12 TEMA E

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6) Shell side inlet nozzle location: Normally it is assumed to be at the tube channel end.

Distance from tangent point to last baffle: Generally the last baffle is placed at the at the tangent point of the U-tube

7) Tube arrangement

8) Impingement device: An impingement plate will be added if the inlet nozzle pV² value is greater than TEMA standards (1500 lb/ft s² or 2232 kg/m s²)

9) Adding a shell in parallel: If pressure drop limitations are not met at the maximum shell diameter permitted, a unit can be added in parallel.

10) Adding a shell in series: If a pure counter flow exchanger has been selected, a shell can be added in series only if the required heat duty can not be achieved at the maximum permitted shell diameter. For multi-tube pass exchangers shells in series can be added if the F-factor is less than 0.7

11) Temperature cross: When the outlet temp of the cold fluid is higher than the outlet temp of hot fluid, it is called the temperature cross. For small exchangers, temperature cross does not have much effect on the type of shells. But for large exchangers shells in series to be used.

12) Baffle design: Baffle window cut should be between 17% to 35%. Baffle spacing should be 20% to 100% of shell ID. For no-tube-in-window exchangers, the ratio of window velocity to cross-flow velocity should normally be 2 to 3. For double segmental baffles (for low pres drop service), baffle spacing should not be too small to avoid ineffective shell side flow patterns.

13) No-tube-in-windows baffle cut design: The baffle cut shall be limited between 15 - 30% of shell diameter

14) Baffle cut out of window: The portion of the baffle which is continuous into the window acts as a sealing strip in the window to force the fluid into the bundle. Continuous baffles should be considered for pull-through floating heads.

Code Y, shell inlet at tube channel end

Code N, shell inlet at tube end Flow 30° Pitch Triangular Pitch 45° Pitch Rotated Square Pitch 60° Pitch Rotated Triangular Pitch 90° Pitch Square Pitch

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12.4 HTRI SHELL SIDE FLOW FRACTIONS

This is for shells with single phase fluid flow. The shell flow is broken down into 5 major streams.

1) B-stream : Main cross flow stream through the bundle. B should be at least 60% of the total flow for turbulent flow and 40% for laminar flow. If the baffle spacing is too narrow, more flow will be forced into the A, C and E streams, thereby decreasing the heat transfer.

2) C-stream : Bundle to shell cross flow bypass stream. C should not normally exceed 10%. Additional sealing strips can be incorporated to decrease this flow fraction. Although this stream is partially effective for heat transfer, a high C-stream flow fraction, especially for pure cross flow shells ("X"), can lead to a severe delta correction to the mean temp difference

3) F-stream : Tube pass partition bypass stream. F should not normally exceed 10%. Additional seal rods can be incorporated to decrease this flow fraction. Although this stream is partially effective for heat transfer, a high C-stream flow fraction, is not recommended. To block the F-stream flow fraction, program assumes one seal rod of a diameter equal to the tube diameter for each 6 tube rows of cross flow in the exchanger

4) A-stream : Tube-to-baffle hole leakage stream. A-stream is large in narrow baffle spacing where large TEMA clearances apply. However, the A-stream is fairly effective thermally. It will decrease for multi-segmental baffles. Fouling layers might seal this A-stream. The design should be examined by giving a zero tube-to-baffle clearance and the built-up fouling layer thickness for a safe design from a pres drop stand point.

5) E-stream : Baffle-to-shell leakage stream. E-stream is highly ineffective thermally because it does not contact the heat transfer surface; but, since it mixes poorly with the other streams, it can cause distortions of the temp profile. If E-stream is more than 15%, double segmental baffle or other modifications should be tested. If E-stream causes (ST program only) low delta correction factor (< 0.8), corrective action is required.

• F-stream seal rods: Allocate one seal rod of the tube diameter for each 6 rows in cross flow in the exchanger

• Sealing strips: These are metal strip or rod placed between the shell and the bundle which has the effect of forcing the bundle bypass C-stream back into the bundle.

Window Baffle A B E A C E A A F C C

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12.5 SHELL SIDE ANNULAR DISTRIBUTOR

Thermal correction factor "F" = (TUBE) x (BAFFLES) x (F/G) x (HOT/COLD); where TUBE is the uncorrected F-factor based on the no of tube passes, shell style and tempBAFFLES is the correction when there are few baffles. (F/G) is the correction for thermal leakage through longitudinal baffle for TEMA "F","G","H" shells.

(HOT/COLD) is the correction for nonconstant overall h.t.coeff due to diff in the h.t.coeff at the hot and cold ends.

Effective MTD = (LMTD) x F x (DELTA): where DELTA is the profile distortion due to the E- and C-stream leakage

12.6 SHELL SIDE HEAT TRANSFER LIMIT

If there is spare shell side pres drop available, the shell side coefficient can be increased by various methods:

1) Changing the shell type to "F" or "G" can increase the shell side velocity and h.t.coefficient. But the mean temperature difference may increase.

2) Reducing the tube pitch

3) Decreasing the tube size to accommodate more tubes in a smaller shell 4) Considering finned tubes

5) If DELTA is lower than 0.85, adding a shell in series gives best result, or using sealing strips might improve the performance

12.7 TUBE SIDE HEAT TRANSFER LIMIT

If there is spare tube side pressure drop available, the tube side coefficient can be increased by various methods:

1) Changing the tube length 2) Decreasing tube diameter

3) When in laminar flow switching the tube side fluid to shell side usually results in a more efficient design

4) Increasing the tube pitch, gives less tubes in the given shell ID

Length

Clearance

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12.8 LIQUID DRIVING HEAD FOR THERMOSIPHON EXCHANGERS

1) For horizontal thermosiphon reboilers, most recirculating type feed systems can be designed with kettle type since the height of the outlet piping entering the column is above the liquid level in the column as shown in figure.

2) For thermosiphon reboiler systems for which the reboiler outlet piping enters the distillation column at a height below the liquid level in the trap-out tray as shown in figurethe piping should be checked to ensure that the liquid level does not cover the exit nozzle.

3) For horizontal thermosiphon reboiler designs, the reboiler exit weight fraction should be limited to 0.5 to avoid tube wall dry out.

4) For an effective design, most of the available loop pressure drop is used across the reboiler. As a rule of thumb, this should be around 60 - 70%.However, the inlet and outlet piping design may change this requirement.

Liquid Driving Head (Static Head) H G D F B E A C C Nozzle pipe Main pipe Main pipe Header pipe Vapor+Liquid

Recirculating Feed System

Vapor+ Liquid

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13. COLUMNS & TOWERS

13.1 GENERAL DESIGN RULES

1) For ideal mixtures, relative volatility can be taken as the ratio of pure component vapor pressures

2) Tower operating pressure is most often determined by the cooling medium in condenser or the maximum allowable reboiler temperature to avoid degradation of the process fluid

3) For sequencing columns:

• Perform the easiest separation first (least trays and lowest reflux)

• If relative volatility or feed composition varies widely, take products off one at time as

the overhead

• If the relative volatility of components does vary significantly, remove products in order

of decreasing volatility

• If the concentrations of the feed vary significantly but the relative volatility does not, remove products in order of decreasing concentration.

4) The three most common types of trays are valve, sieve, and bubble cap. Bubble cap trays are typically used when low-turn down is expected or a lower pressure drop than the valve or sieve trays can provide is necessary

5) Bubble cap trays are used only when a liquid level must be maintained at low turn down ratio; they can be designed for lower pressure drop than either sieve or valve trays.

6) The optimum Kremser absorption factor is usually in the range of 1.25 to 2.00

13.2 TRAY COLUMNS/TOWERS

1) The most economic number of trays is usually about twice the minimum number of trays 2) The minimum number of trays is determined with the Fenske-Underwood Equation 3) Typically, 10% more trays than calculated are specified for a tower

4) Tray spacing’s should be from 18” to 24” (450 to 600 mm), with accessibility in mind

5) Peak tray efficiencies usually occur at linear vapor velocities of 2 ft/s (0.6 m/s) at moderate pressures, or 6 ft/s (1.8 m/s) under vacuum conditions.

6) Pressure drop per tray is of the order of 3” of water or 0.1 psi (0.007 bar)

7) Tray efficiencies for distillation of light hydrocarbons and aqueous solutions are usually in the range of 60-90% while gas absorption and stripping typically have efficiencies closer to 10-20%

8) Sieve tray holes are 0.25 to 0.50 in. diameter with the total hole area being about 10% of the total active tray area

9) Valve trays typically have 1.5” diameter holes each with a lifting cap. 12-14 caps/square foot of tray is a good benchmark. Valve trays usually cost less than sieve trays

10) The most common weir heights are 2” and 3” (50 to 80 mm) and the weir length is typically 75% of the tray diameter

11) For towers that are at least 3 ft (0.9 m) in diameter, 4 ft (1.2 m) should be added to the top for vapor release and 6 ft (1.8 m) should be added to the bottom to account for the liquid level and reboiler return

12) Limit tower heights to 175 ft (53 m) due to wind load and foundation considerations 13) The Length/Diameter ratio of a tower should be no more than 30 and preferably below 20 14) Liquid redistributors are needed every 5-10 tower diameters with pall rings but at least

every 20 feet (6.1 m). The number of liquid streams should be 3-5 /ft² in towers larger than 3 ft dia, and more numerous in smaller towers

15) A rough estimate of reboiler duty as a function of tower diameter is given by: Q = 0.5 D² for pressure distillation

Q = 0.3 D² for atmospheric distillation Q = 0.15 D² for vacuum distillation

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Where Q = Reboiler duty (Million Btu/h) D = Tower diameter (ft)

13.3 REFLUX DRUMS

1) The most economic reflux ratio usually is between 1.2 Rmin and 1.5 Rmin 2) Reflux pumps should be at least 25% over designed

3) Reflux drums are almost always horizontally mounted and designed for a 5 min holdup at half of the drum's capacity. A take off pot for second liquid phase such as water in hydrocarbon systems, is sized for a linear velocity of that phase of 0.5 ft/s (0.15 m/s) with minimum diameter of 16” (400 mm)

13.4 PACKED TOWERS

1) Packed towers almost always have lower pressure drop than comparable tray towers 2) Packing is often retrofitted into existing tray towers to increase capacity or separation 3) For gas flow rates of 500 ft³/min (14.2 m³/min) use 1 in (2.5 cm) packing, for gas flows of

2000 ft³/min (56.6 m³/min) or more, use 2 in (5 cm) packing

4) Ratio of tower diameter to packing diameter should usually be at least 15

5) Due to the possibility of deformation, plastic packing should be limited to an unsupported depth of 10-15 ft (3-4 m) while metallic packing can withstand 20-25 ft (6-7.6 m)

6) Liquid distributor should be placed every 5-10 tower diameters (along the length) for pall rings and every 20 ft (6.5 m) for other types of random packing

7) For redistribution, there should be 8-12 streams per sq. foot of tower area for tower larger than 3 feet in diameter. They should be even more numerous in smaller towers

8) Packed columns should operate near 70% flooding

9) Height Equivalent to Theoretical Stage (HETS) for vapor-liquid contacting is 1.3-1.8 ft (0.4-0.56 m) for 1 in pall rings and 2.5-3.0 ft (0.76-0.90 m) for 2 in pall rings

13.5 DESIGN PRESSURE DROPS

Service Pressure drop (in H2O/ft packing)

Absorbers and Regenerators

Non-Foaming Systems 0.25 - 0.40

Moderate Foaming Systems 0.15 - 0.25

Fume Scrubbers

Water Absorbent 0.40 - 0.60

Chemical Absorbent 0.25 - 0.40

Atmospheric or Pressure Distillation 0.40 - 0.80

Vacuum Distillation 0.15 - 0.40

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14. COMPRESSORS AND VACUUM EQUIPMENT

The following chart is used to select the type of compressor:

1) Fans should be used to raise pressure about 3% (12” or 300 mm water), blowers to raise to less than 2.75 barg (40 psig), and compressors to higher pressures

2) The theoretical reversible adiabatic power is estimated by:

[

]

a a P P T R Z m Power 1 ( 2/ 1) 1 − × × ⋅ ⋅ ⋅ = where:

T1 = Inlet temperature; R = gas constant; Z1 = Compressibility; m = molar flow rate

(

)

a P P T T Cv Cp k k k a × × = = − = ) / ( / / 1 1 2 1 2

T2 = Outlet temperature for adiabatic reversible flow 3) Exit temperatures should not exceed 204°C (400°F)

4) For diatomic gases (Cp/Cv = 1.4) this corresponds to a compression ratio of about 4 5) Compression ratios should be about the same in each stage for a multistage unit,

n n P P Ratio / 1 1⎟⎟⎠ ⎞ ⎜⎜ ⎝ ⎛ = with n stages.

6) Efficiencies for reciprocating compressors are:

- 65% at compression ratios of 1.5

- 75% at compression ratios of 2.0

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7) Efficiencies of large centrifugal compressors handling 2.8 to 47 m³/s (6000-100,000 acfm) at suction is about 76-78%

8) Flash gas compressors typically have an overall compressor ratio in the range 5 to 10 9) For compressors, the brake horsepower per stage can be determined from

( ) ( )

⎥ ⎥ ⎥ ⎦ ⎤ ⎢ ⎢ ⎢ ⎣ ⎡ − ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ × ⎟ ⎠ ⎞ ⎜ ⎝ ⎛ − ⋅ ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ ⋅ ⋅ ⋅ ⋅ = − − 1 1 0857 . 0 1 1 1 k k s d s g k k s k av P P k k E T Q Z Z BHP

Where BHP = Brake horsepower per stage Qg = Volume of gas (MMscfd) Ts = Suction temperature (R) Zs = Suction compressibility factor Zd = Discharge compressibility factor

Zav = (Zs + Zd) / 2 E = Efficiency %

High speed reciprocating units – use 70% Low speed reciprocating units – use 78% Centrifugal units - use 75% K = Ratio of specific heat, Cp/Cv

Ps = Suction pressure of stage (psia) Pd = Discharge pressure of stage (psia)

14.1 VACUUM PUMPS

1) Reciprocating piston vacuum pumps are generally capable of vacuum to 1 torr (1 mmHg absolute)

2) Rotary piston types can achieve vacuums of 0.001 torr

3) Single stage Jet ejectors are capable of vacuums up to 100 torr, 2-stages to 10 torr, 3-stages to 1 torr, and 5-3-stages to 0.05 torr

4) A three stage ejector requires about 100 lb steam/lb air (100 kg steam/kg air) to maintain a pressure of 1 torr

5) Air leakage into vacuum equipment can be approximated as follows: 3 / 2 V k Leakage= ×

Where k = 0.20 for P >90 torr, 0.08 for 3 < P < 20 torr, and 0.025 for P < 1 torr V = equipment volume (ft³)

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15. SAFETY SYSTEM & PSV DESIGN

15.1 SAFETY SYSTEM CONFIGURATION

1) For direct discharge of fluids to atmosphere, they should be in the vapor state, below their auto ignition temperature, and should meet one of the following requirements:-

• Flammable vapors of MW less than 28.9 (MW of air).

• Flammable vapors heavier than air with MW less than 70 but with minimum discharge velocity of 500 fps (152 m/s), based on maximum capacity of the relief valve

• Vapors of any MW that are non-flammable, non-toxic and non-condensable 2) In all other cases relieved fluids should be disposed to a flare system.

3) Boilers having more than 47 m² of water heating surface or electric boilers having a power input of more than 500 kW shall have 2 or more PSV.

4) For electric boilers the minimum relieving capacity shall be 1.6 kg/hr/kW input

5) For Boilers, if additional PSV are used the highest pressure setting shall not exceed the MAWP by more than 3%. When multiple PSV are installed, the difference between the highest and lowest set pressure should not be greater than 10% of the highest set pressure.

6) For Pressure vessels, if additional PSV are used the highest pressure setting shall not exceed 105% of MAWP. For multiple PSV, one of which is installed for fire exposure only, this particular valve may be set at a pressure not exceeding 110% of MAWP

7) Fire exposure protection of vessels by water spray @ 0.05 - 0.2 gpm/ft² of total vessel area. Fire exposure is considered for a wetted area up to 25 ft (7.62 m) from grade. Average or normal liquid level to be considered for vessels, High liquid level for columns, surge drums and KO drums. 50% of storage tank height or 25 ft (7.62 m) from grade, which ever is higher.

8) For flare fire relief load, consider plot area between 2000-5000 ft² (186-465 m²). Use 2500 ft² (232 m²) for a paved drained surface in a plant where NFPA required fire fighting equipment is available.

9) If fire proofing insulation is not provided, environmental factor (F) should be = 1.0

15.2 SAFETY VALVE BACK PRESSURE

1) Super imposed back pressure = pressure at PSV discharge before valve opening 2) Builtup back pressure = pressure at PSV discharge header after valve opening.

3) Conventional PSV shall be used when the superimposed back pressure varies over a range not exceeding 10% of the set pressure (gauge). However the performance of PSV at builtup back pressure to be studied by the vendor before selecting the PSV

4) Balanced type PSV can be used for >10% superimposed back pressure. However when builtup back pressure exceeds 30-50% of set pressure, capacity of PSV for vapors and gases starts to fall below the theoretical capacity. With liquids, the capacity reduction starts at 15% of set pressure. The fall in capacity depends on overpressure, type and make of PSV.

5) Pilot operated relief valves are used when operating pressure is very close to set pressure or when set pressure is below 10-15 psig (0.7 – 1.0 barg)

15.3 SAFETY VALVE RELIEF LOAD CALCULATION

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1) Blocked outlet 2) Gas blow by case

3) Inadvertent valve opening 4) Reflux failure

5) Tube rupture 6) Fire case 7) Reverse flow 8) Thermal relief case.

15.3.1 Blocked Outlet

Blocked outlet case can occur when the control valves at the outlet of the vessel close at the same time due to the stoppage of instrument air supply or plant shut down.

The relief quantity shall be the design flow rate to the vessel.

15.3.2 Fire Case

The amount of heat absorbed by a vessel open to fire is markedly affected by the type of fuel feeding the fire.

The following equivalent formulas are used to evaluate the condition where there are prompt firefighting efforts and drainage of flammable materials away from the vessels are available:

18 . 0 21000× × − = F A q 82 . 0 21000 F A Q= × ×

Where adequate drainage & firefighting equipment do not exist, the following equation should be used: 82 . 0 34500 F A Q= × ×

Where q = Average unit heat absorption (Btu/h/ft² of wetted surface) Q = Total heat absorption (input) to the wetted surface (Btu/h)

F = Environmental factor (Values for various type of installation are given in Table 5 in API 521 page 25)

A = Total wetted surface (ft²)

The discharge area for pressure relief devices on vessels containing super critical fluids, gases or vapors exposed to open fires can be estimated by using the following equation.

1 ' ' P A F A= × ⎟ ⎟ ⎠ ⎞ ⎜ ⎜ ⎝ ⎛ − ⋅ = ₀_.₆₅₀₆ 1 25 . 1 1) ( 1406 . 0 ' T T T K C F w D

The relief load can calculated directly, in pounds per hour.

⎟ ⎟ ⎠ ⎞ ⎜ ⎜ ⎝ ⎛ − × ⋅ × = ₁_.₁₅₀₆ 1 25 . 1 1 1 ) ( ' 1406 . 0 T T T A P M W w