Process Equipment Design (Pinoy Edition)

Chemical Engineering Handbook Process Equipment Design

O. PROCESS EQUIPMENT DESIGN

O.1. PIPE AND TUBE SIZING

The choice of pipe diameter can significantly affect the operating cost and capital investment requirement in the operation of an industrial plant. Piping including installation could go up to 17% of the Fixed Capital Investment of a chemical processing plant (Peters et al, 2004). The choice of tube diameter would dictate the tube fluid velocity of cooling water in heat exchangers. At high cooling water velocity, convective heat transfer coefficient increases which would eventually result to smaller heat transfer area requirement hence lower equipment cost. However, at high fluid velocity, pumping cost would also be higher, hence higher operating cost. Therefore a good compromise has to be arrived in the selection of an appropriate pipe and tube diameters.

Kent (1980) derived typical pipe diameter based on typical fluid velocity. Minimum pipe diameter was also derived based on maximum velocity. In this scheme, fluids (liquids and gases) are classified into clean and erosive/corrosive fluids.

O.1.1. Typical Pipe Diameter

Typical diameter based on typical fluid velocity

for liquids

(O – 1)

for gases

(O – 2)

Minimum diameter based on maximum fluid velocity

Clean Fluids

for liquids

(O – 3)

O - 1


O - 2


for gases

(O – 4)

Erosive or Corrosive Fluids

For liquids

(O – 5)

for gases

(O – 6)

where: D = Pipe inside diameter, inch

W = Flow rate in 1000

= Density,

m = Molecular weight

K = Piping Cost,

Z = Compressibility factorT = Rankine temperature

O.1.2. Economic Pipe Diameter

Peters et al (2004), presented an estimation of economic pipe diameter considering pipe, installation, fittings and maintenance costs. The applicable estimations have been categorized according to the type of flow regime. Under each flow regime, a subcategory has been established according to the size range of the pipe diameters.

Turbulent Flow in Steel Pipes

for Di 1 inch

(O – 7)

O - 3


for Di < 1 inch

(O – 8)

For viscosity range of 0.02 to 20 centipoises, and µ 0.025 maybe taken as unity.

Viscous Flow in Steel Pipes

Di 1 inch

(O – 9)

Di < 1 inch

(O – 10)

Note: Diopt is in meter if constants inside the bracket of Equations O – 7 to O – 10 are replace with 1.63 x 10 -6, 1.53 x 10-5, 4.39 x 10-4, and 4.14 x 10-3 respectively and units inside the parenthesis are used.

where Di = optimum pipe diameter, in. (m)

qf = Volumetric flowrate, (m3/s)

= Fluid viscosity, centipoises

K = Cost of electrical energy,

J = Frictional loss due to fittings and bends, expressed as equivalent frictional loss in a straight pipe

Hy = Hours of operation per yearF = Ratio of Total Cost for fittings and installation to purchase cost for new

PipeE = Efficiency of motor and pump expressed as a fractionKF = Annual fixed charges including maintenance, expressed as a fraction of

initial cost for completely installed pipeX = Purchase cost of new pipe per foot of pipe length if pipe diameter is

one inch (0.0254 m),

Other alternative equation in determining optimum economic pipe diameter has been presented as:

O - 4


Incorporating pump power Kent (1980), recommends the equation below:

(O – 11)

where C = Pump Cost per Actual horsepower,

For Di , incorporating cost of Capital or Return on Investment, cost of pump, taxes and time value of money, use eqn. 49 (page 366) or 9-80 (405) of Peters and Timmerhaus (1991) or (2004) as the case maybe.

O.1.3. Pipe and Tube Diameters Based on Velocity

Pipe and Tube diameters may be also computed using acceptable velocity at a known volumetric flow rate. With this approach, the continuity equation provides for the computation of pipe or tube diameters. Typical fluid velocity resulting from the design and operational specifications will give a typical pipe diameter. While economic velocities resulting from the consideration of design and operational economics may be also used to estimate optimum economic pipe diameter.

Backhurst and Harker (1973) presented reasonable liquid and gas velocities through the tubes to be 3 – 5 ft/s and 50 – 100 ft/s respectively. Another estimation (Perry and Green, 1997) of typical economic optimum velocities for low viscosity liquids in schedule 40 steel pipe have been provided to be 5.9 – 7.9 ft/s and 30 to 131 ft/s for gases with density ranging 0.013 to 1.25 lb/ft3.

In general, an optimum liquid velocity ranges from 5 – 10 ft/sec (Baasel, 1974). Thermodynamic considerations favor small tube diameters for a compact heat exchanger. However, tube cleaning practices limit the tube diameter to a minimum of approximately 20 mm outer diameter (Hewitt, et. al., 1994) ¾” and 1” OD are the most common sizes while 5/8 to 1 ½ inch are also found in industrial applications (Peters and Timmerhaus, 1991). On the other hand, allowable maximum fluid velocity may be varied depending on the tube material resistance to corrosion or vibration damage. A power plant condenser with sea water flowing is limited to 2 m/s with Cu-Ni tubes, but could operate at up to 4 m/s with titanium tubes. Higher cost of tubing (titanium) is hereby offset by high overall heat transfer coefficient value (resulting to a smaller unit), and virtual absence to corrosion and hence no need for retubing (Hewitt et. al, 1994).

O - 5


More detailed typical velocities have been provided in Tables O – 1 to O – 3 to estimate economic pipe diameter (Kent, 1974). While similar detailed economic velocities have been provided in Table O – 4 to estimate the same (Peters and Timmerhaus, 1991).Table O-1. Typical Liquid Velocities in Steel Pipelines.

Nominal Pipe Sizes, In. 2 or less 3 to 10 10 to 20Liquid and line Velocity, Ft/s Velocity, Ft/s Velocity, Ft/s

WaterPump Suction 1 to 2 2 to 4 3 to 6Pump discharge (long) 2 to 3 3 to 5 4 to 7Discharge leads (short) 4 to 9 5 to 12 8 to 14Boiler feed 4 to 9 5 to 12 8 to 14Drains 3 to 4 3 to 5 ---Sloped sewer --- 3 to 5 4 to 7

Hydrocarbon liquids(Normal viscosities)Pump suction 1.5 to 2.5 2 to 4 3 to 6Discharge header (long) 2.5 to 3.5 3 to 5 4 to 7Discharge leads (short) 4 to 9 5 to 12 8 to 15Drains 3 to 4 3 to 5 ---

Viscous oilsPump suction

Medium viscosity --- 1.5 to 3 2.5 to 5Tar and fuel oils --- 0.4 to 0.75 0.5 to 1

Discharge (short) --- 3 to 5 4 to 6Drains 1 1.5 to 3 ---

Adapted from Table 4 p 89. Process Piping Systems. (Deutsch 1980)

Table O – 2. Typical Velocities in Gas and Vapor Lines.Nominal Pipe

Size, In.Saturated Steam or

Saturated VaporSuperheated Steam,

Superheated Vapor or GasLow Pressure Medium Pressure High PressureVelocity, Ft/s Velocity, Ft/s Velocity, Ft/s

2 or less 45 to 100 40 to 80 30 to 603 to 4 50 to 110 45 to 90 35 to 70

6 60 to 120 50 to 120 45 to 908 to 10 65 to 125 80 to 160 65 to 12512 to 14 70 to 130 100 to 190 80 to 14516 to 18 75 to 135 110 to 210 90 to 160

20 80 to 140 120 to 220 100 to 170Adapted from Table 3 p 89. Process Piping Systems. (Deutsch 1980)Note: Within the above velocities and line-size ranges, (a) larges lines can have higher velocities than smaller ones, and (b) short lines, and leads from headers, can have higher velocities than long lines and headers.

Table O – 3. Typical Velocities in Equipment Lines.Equipment Lines Velocity, Ft/sReboiler, downcomer (liquid) 3 to 7

O - 6


Reboiler, riser (liquid and vapor) 35 to 45Overhead condenser 25 to 100

Adapted from Table 4 p 89. Process Piping Systems. (Deutsch 1980)

Table O – 4. “Rule-of-Thumb” Economic Velocities for Sizing Steel Pipelines.

Turbulent flowType of Fluid Reasonable velocity, ft/sWater or fluid similar to water 3-10Low-pressure steam (25psig) 50-100High-pressure steam (100psig and up) 100-200Air at ordinary pressures (25-50 psig) 50-100

The preceding values apply for motor drives. Multiply indicated velocities by 0.6 to give reasonable velocities when steam turbine drives are used.

Viscous Flow (liquids)Nominal pipe diameter, in.

Reasonable velocity, ft/sc = 50 c = 100 c = 1000

1 1.5-3.0 1.0-2.0 0.3-0.62 2.5-3.5 1.5-2.5 0.5-0.84 3.5-5.0 2.5-3.5 0.8-1.28 4.0-5.0 1.3-1.8

c = viscosity, centipoiseAdapted from Table 2, p 496. Plant Design and Economics for Chemical Engineers 4th ed.

O.2. PIPE AND TUBE THICKNESS

In the following paragraphs are the classical presentations reproduced from Hesse and Rushton (1975). In accordance with ASME Code, the maximum allowable internal working pressure for Ferrous tubes and pipes is given by:

(O – 12)

Rearranging, the above equation provides for the wall thickness as:

(O – 13)

where t = pipe thickness, inchD = outside pipe diameter, inchP =working pressureSF = allowable tensile strength for Ferrous Material (Table O - 5)

For Non-Ferrous tubes and pipes, the following equation is given:

O - 7


(O – 14)

SNF = allowable tensile strength for Non-Ferrous Material (Table O - 6)

The above equations are only applicable on the following conditions:

1. Outer diameters should be between 0.5 and 6 inches2. Wall thickness should not be less than 0.049 inch

Furthermore, an additional wall thickness must be provided when corrosion or wear due to cleaning is to be anticipated. Tubes that are threaded should provide for a wall allowance of

. Also, if the tubes are rolled into headers, additional wall

allowance to compensate for the thickness reduction due to rolling should be incorporated.

Standard number of Thread per Linear Inch (Hesse & Rushton, 1945)Nominal Diameter

1/8 ¼ 3/8 ½ & ¾1, 1¼, 1½,

22½ - 10

No. of Threads per Inch

27 18 18 14 11½ 8

For steam piping application, minimum thickness is given by

(O – 15)

where corrosion allowance C could have any of the following values:

C = 0.05 – for inch and smaller threaded pipe and for plain end pipe or tubing 1

inch nominal size and smaller

C = – for ½ inch and larger threaded pipe

C = 0.065 – for plain end pipe over 1 inch nominal size

The above equation for steam piping is applicable for steel or wrought iron pipe whose nominal size is 4 inches or less. For nominal pipe diameter larger than 4 inches, the applicable equation is:

O - 8


(O – 16)

It should be noted that for steam pressures greater than 250 psi and for water pressure and temperature greater than 100 psi and 200ºF, seamless pipe of quality equivalent to S-17 or S-18 and of a weight equivalent to schedule 80 is the minimum requirement.

Working equations used for steam piping may be extended for oil piping use, provided the allowable stress should be multiplied by 1.25, since all allowable stresses for petroleum liquids and gases are based on apparent factor of safety of 4 (API-ASME), instead of 5 as recommended in ASME-PB and ASME-UPV Codes.

O.2.1. Cast Iron Pipe

Cast iron pipes with flanged ends can be used for gas and oil service for underground application where the metal temperature of the pipeline is less then 300F. Aboveground, it may be used for pressures not greater than 150 psi, where metal temperatures do not exceed 300ºF. The working equations used for steam piping are also applicable for cast iron wall thickness calculation. However, an allowance to internal pressure is added to incorporate water hammering effect as follows:

Nominal Pipe Size Added Pressure (psi)

4 – 10” 12012 – 14” 11016 – 18” 10042 – 60” 70

Also, the allowable stress (S) and allowance (C) may be taken as:

S = 4,000 psi for pipe cast vertically in dry sand moldS = 6,000 psi for pipe cast centrifugally or horizontallyC = 0.18 for vertical or pit cast pipeC = 0.14 for centrifugally cast pipe

O - 9


Table O - 5. Allowable Stresses, psi, for Ferrous Materials for Pipes and Tubes. (ASME-UPV Code)

Spec. ASTM For Temperature not Exceeding Degrees FNo. Grade Spec. Weld 650 700 750 800 850 900 950

S-17 Steel A83-38T Lap 7,300 7,000 6,650

S-17 Steel A83-38T Seamless 9,400 9,000 8,150 7,150 5,850 4,400 2,600

S-17 Wrought Iron A83 Lap 5,600 5,300 4,800

S-18 Steel A53-36 Lap 7,300 7,000 6,650

S-18 Steel A53-36 Butt 5,400 5,300 5,050

S-18 Steel A53-36 Seamless 9,600 9,100 8,250 7,250 5,850 4,400 2,600

S-19 Wrought Iron A72-39 Lap 5,600 5,300 4,800

S-19 Wrought Iron A72-39 Butt 4,800 4,600 4,150

S-32 A, Silicon 0.10% A178-37 Resistance 8,000 7,650 7,300 6,700 5,800 4,750 3,200

S-32 A A178-37 Resistance 8,000 7,650 6,950 6,100 4,950 3,750 2,200

S-32 B A178-37 Resistance 6,800 6,500 5,850

S-32 C, Silicon 0.10% A178-37 Resistance 10,200 9,700 8,850 7,750 6,300 4,750 3,250

S-32 C A178-37 Resistance 10,200 9,700 8,450 7,050 5,400 3,750 2,200

S-34 P3a A158-38T Seamless 12,000 12,000 12,000 11,800 11,200 10,000 8,000

S-40 A A192-38T Seamless 9,400 9,000 8,600 7,900 6,800 5,600 3,800

S-45 P1 A206-39T Seamless 11,000 11,000 11,000 10,700 10,500 10,000 8,000

S-49 A210-38T Seamless 12,000 11,400 10,400 9,100 7,400 5,600 3,800

S-48 T1 A209-38T Seamless 11,000 11,000 11,000 10,750 10,500 10,000 8,000

S-48 T1a A209-38T Seamless 12,000 12,000 12,000 11,500 11,000 10,000 8,000

Table O - 6. Allowable Stresses, psi, for Non-Ferrous Materials for Pipes and Tubes. (ASME – UPV Code)

For Metal Temperature Not Exceeding Degrees FSpec. Subzero

Material No. to 150 250 300 350 400 450 500 550 600

S-24Muntz metal S-47 10,000 9,000 5,500 2,000 1,500

S-59

Red Brass, high brass S-24 7,000 6,500 5,750 5,000 3,000 1,000 800

S-24Admiralty S-47 7,000 6,500 6,250 6,000 5,500 4,500

Naval brass * 11,000 10,000 10,000 6,500 3,000

Steam bronze S-41 7,000 7,000 6,500 6,000 5,500 5,000 4,000 3,000

Steam bronze S-46 6,000 5,500 5,000 4,500 3,500

O - 10


Monel metal† S-54 14,000 14,000 14,000 14,000 14,000 14,000 14,000 14,000 14,000

Cupro-nickel 70-30† S-47 11,000 11,000 11,000 11,000 11,000 11,000 10,000 10,000 9,000

Cupro-nickel 80-20† S-47 10,000 10,000 10,000 10,000 10,000 10,000 9,000 9,000 8,000

S-20S-22

Copper, annealed, all S-23 6,000 5,000 4,750 4,500 4,000

types S-47S-66

Aluminum manganese S-39 2,800 2,400 2,100 1,800 1,600

alloy, annealed

Aluminum manganese

alloy, quarter-hard or S-39 3,500 3,000 2,700 2,400 2,200

as rolled

* U.S. Navy Dept. Spec. 46B-6-j†Maximum permissible temperature Monel metal and cupro-nickel 750°F

O.2.2. For Other Pipe Materials of Construction

Working equations for pipe wall thickness for different materials may be derived from different Philippine National Standard (PNS).

Unplasticized Polyvinyl Chloride (PVC) Pipe for Potable Water Supply (PNS 65,19)

(O – 17)

High Density Polyethylene (PE) pipe for Potable Water Supply (PNS 55, 19)

(O – 18)

where

(O – 19)

O - 11


where = standard thermoplastic pipe dimension ratio

Polybutylene (PB) pipes for Potable Water Supply (PNS 152, 22)

(O – 20)

(O – 21)

Machine – Made Filament Wound Fiberglass – Reinforced Thermosetting Resin Pipes (PNS 128, 21)

(O – 22)

where S = hydrostatic design stress, MpaP = Pressure rating, MPaD = average outside diameter, mmT = minimum wall thickness, mm

O.3. STORAGE TANK DESIGN

Storage tanks have been an integral part of many industrial plants and processes. Design of tanks under consideration covers tank sizing, shell thickness, stiffers spacing and annular plate calculations. This section is limited for liquid storage only.

O.3.1. Tank Sizing

Heuristics or Rules of thumb will be used for quick tanks sizing. Tanks may be classified as Field erected or Shop – constructed. Tank capacities greater than 23,700 gallons (90 m3) should be constructed to the dimensions given by the American Petroleum Institute (API) Standards (Baasel, 1974). Although, these rules will not apply to all tank sizing, they could be very useful in many circumstances.

For Shop – constructed tanks, optimum tank diameter (Mac Cary, 1960) is calculated using:

(O – 22)

where D = diameter in ft (m)V = volume in ft3 (m3)

For Field – erected tanks, API typical sizes is given in Table O – 5.

O - 12


Table O – 7. Selection of Typical Sizes of API Field Constructed Tanks.

Diameter Approx. Capacity Height Volumeft m Gal/ft m3/m ft m gal m3

15 4.6 1320 16.4 18 5.5 23,700 9020 6.1 2350 28.0 18 5.5 42,500 16125 7.6 3670 45.6 18 5.5 66,000 25025 7.6 3670 45.6 24 7.3 88,000 33430 9.1 5290 65.6 24 7.3 127,000 48135 10.7 7190 89.3 30 9.1 216,000 81945 13.7 11900 148.0 36 11.0 429,000 162570 21.3 28800 358.0 36 11.0 1040,000 3940100 30.5 58700 728.0 36 11.0 2110,000 8000120 36.6 84500 1050.0 48 14.6 4060,000 15400180 54.9 190000 2380.0 48 14.6 9150,000 34700

SOURCE: “Welded Steel Tanks for Oil Storage,” American Petroleum Institute, Washington, D.C. 1973.

The selection of shop – constructed fabrication would prove to be less expensive, however, tanks constructed under this category should not be more than 11.5 ft (3.5 m) in diameter due to transport limitation (Baasel, 1974).

Furthermore guidelines on allowances, tank orientation and mounting have been given:

1. Vessels below 500 gallons (1:9 m3) are never more than 85% filled2. Vessels above 500 gal (1.9 m3) are never more than 90% filled.3. Liquid in quantities less than 1,000 gal (3.8 m3) is stored in vertical tanks mounted on legs.4. Liquid in quantities between 1,000 and 10,000 gal (3.8 and 38 m3) is stored in horizontal

tanks mounted on a concrete foundation.5. Liquid in quantities exceeding 10,000 gal (3.8 m3) is stored in vertical tanks mounted on a

concrete foundation.

O.3.2. Shell Thickness

One of the standards used to design flat – bottom tanks is that of American Petroleum Institute (API). API 650 is used to design welded steel tanks for oil storage. The following paragraphs have been reproduced from API 650 for flat – bottom tanks containing liquids with little or no surface pressure as cited by Jawad and Farr (1988).

Flat bottom tanks are normally constructed according to one of the following four standards:

I. API 650. American Petroleum Institute Welded Steel Tanks for Oil Storage.

II. API 620. Recommended Rules for Design and Construction of Large Welded Low Pressure Storage Tanks.

O - 13


III. ANSI B96.1. American National Standard Institute for Welded Aluminum-Alloy Storage Tanks.

IV. AWWA-D-100. American Water Works Association Standard for Welded Steel Tanks for Water Storage.

A. API 650 Tanks

The requirements of API 650 are for flat-bottom tanks containing liquids with little or no surface pressure. The design criteria are based on simplified equations with a minimum amount of analysis.

1. Roof Design

Flat-bottom tank with large diameter and fixed roof normally are designed with column-supported roofs. As the diameter gets smaller, self-supporting roofs become more economical. Dome and cone roofs are the most popular types.

The following equation for designing self-supporting dome roofs incorporates a factor of safety (FS) of 4.

(O – 23)

The required thickness is obtained by assuming the maximum pressure consisting of a live load of 25 psf, which is the assumed snow load, and a dead load of a maximum roof thickness of 0.5 inch as allowed by API.

Hence:P = 25 psf live load + 20.4 psi dead loadP = 0.315 psi

Letting E = 29 x 106 psi and expressing R (spherical radius of dome roof) in feet and t in inches equation (O - 23) becomes

(O – 24)

which gives the required thickness of a dome roof.

2. Shell Design

O - 14


API 650 includes two methods for the design of shells. The first, called the one-foot method, consists of calculating the required thickness of shell course A in Figure O - 1 based on the hydrostatic pressure at 1 ft above point X, which is the circumferential seam between courses A and B. This method considers that the bottom plate on course B stiffens the next course at point X and the maximum stress occurs at a location higher than X. This location is arbitrarily set at one foot.The required thickness is given as:

(O – 25)

where: CA = Corrosion allowance (in)D = Tank diameter (ft)G = Specific gravity of liquidH = Liquid height (ft)S = Allowable stress (psi)t = Required thickness (inches)

Figure O - 1. Stress Along the Shell Tank Height.

The second method, the variable point method, is an extension of the one-foot method in that it calculates a more exact location of the maximum stress near the junction of the bottom or shell courses with differing thickness. In this case the bottom course is assumed to be hinged at its junction with the bottom plate. Hence, the deflection due to internal pressure at the junction is equal to the deflection due to an applied shearing force (Jawad and Farr, 1988).

The modified equation for the bottom course thickness is given as:

(O – 26)

While the thickness of the second course is determined from the equations below:(O – 27)

O - 15

1’

1’

A

X

BD/2

H


where t1 = thickness of first course (inches)t2 = thickness of second course (inches)

t2a = thickness of second course calculated from the equation for upper course (inches)

h1 = height of first course (inches) r = radius of shell (inches)

Design of the upper courses is based on the equation:

(O – 28)

where: X is a variable design point that is a function of the thickness of layers, tank radius, and liquid height. It is the minimum value of X1, X2 and X3 obtained from the following equations:

where

hu = height of upper shell (inches)tL = thickness of lower shell (inches)tu = thickness of upper shell (inches)

After establishing the shell thickness due to hydrostatic pressure, stability under wind loads must be checked. The applied wind pressure is normally expressed as:

P = 0.00256 V2 (O – 29)

where P = wind pressure (psf)V = Wind velocity (mph)

O - 16


API uses a 100 mph wind velocity for design purposes unless higher value is specified. Hence,

P = 25.6 psf

Because the pressure distribution may cause a vacuum on any part of the shell, the shell is designed to withstand a vacuum pressure of 25.6 psf. A simplified expression for length between stiffeners for the buckling of long cylindrical shells with E = 30 x 106 and Poisson’s ratio of 0.3 and FS = 1.5 is

(O – 30)

where H = length between stiffeners (feet)t = thickness of shell (inches)D = tank diameter (feet)

NOTE: Most cylindrical shells are subjected to various (both external and internal) compressive forces such as dead weight, wind loads, earthquakes and vacuum. The behavior of cylindrical shells under these compressive forces is different from those under internal pressure. Cylindrical shells are weaker in compression that in tension.For maximum strength of cylindrical shells under external pressure, ends of shells are simply supported. For this to be true, stiffening rings, flanges and so on are needed as lines of supports. The supports are assumed to carry all the load that the shell carries due to external pressure.

Table O - 8. Various Standards Requirements for Flat-Bottom Tanks.API 650 API 620

Appendix Appendix Appendix Appendix ANSI

Basic A F Basic R Q B96.1 AWWA

Maximum internal P atm. atm 2.5 psi 15 psi 15 psi 15 psi atm atm.Minimum temperature NS -20°F NS -50°F -60°F -270°F -20°F -55°FMaximum temperature 200°Fa 200°F 200°Fa 200°F -40°F 200°F 400°F RTMaximum shell thickness 1.75 in 0.50 in. 1.75 NS NS NS NS 2 ine

Minimum shell thicknessD < 50 ft 0.1875 in 0.1875 0.1875 inb 0.1875 in

50 ft < D < 120 ft 0.25 in. 0.25 in 0.25 inc 0.25 in120 ft < D < 200 ft 0.3125 in. 0.3125 in 0.3125 in 0.3125 in

D > 200 ft 0.375 in 0.375 in. 0.375 inMinimum roof thickness 0.1875 in. NS 0.1875 in 0.1875 inf

Minimum bottom-platethickness 0.25 in. + CA NSd 0.25 in. NSd

Min. top angleD < 35 ft 2 in. x 2 in. x 3/16 in NS 2½ in. x 2½ in. x ¼ in. NS

35 ft < D < 60 ft 2 in. x 2 in x ¼ in. NS 2½ in x 2½ in. x 5/16 in. NSD > 60 ft 3 in. x 3 in. x 3/8 in. NS 3 in. x 3 in. x 4/8 in. NS

Adapted from “Structural Analysis and Design of Process Equipment (Jawad and Farr,1988).NS = not specified, CA = corrosion allowance, RT = room temperatureaThis temperature can be extended as high as 500°F when certain additional material and design requirements are metbThis thickness applies to tanks with a diameter less than 20 ft.cThis thickness applies to tanks between 20 and 120 ft. in diameterdMinimum thickness of any plate is 0.1875 in. plus CAeFor thicknesses over 2 in., additional requirements must be metfFor cone roofs, the plate thickness may be 7 gauge steel

O - 17


3. Annular Plates

The required thickness of the bottom plate in an API 650 tank is given in Table O - 8. At the shell to bottom plate junction, the API standard requires a butt-welded annular plate whose thickness varies between 0.25 and 0.75 inch and is a function of the stress and thickness of the first shell course. The width of the annular plate must be adequate to support the column of water on top of it in case of a foundation settlement.

Using a FS of two the length of the annular plate is thus expressed as:

(O – 31)

where L = length of annular plate (inches)tb = thickness of annular plate (inches)H = height of liquid (feet)G = specific gravity of liquid

O.4. WELDED PRESSURED VESSEL (UNFIRED)

Unfired Pressured Vessels (UPV) in consideration may include reactors, storage tanks, fractionating column, heat exchangers and so on.

In a classical chemical engineering equipment design class, Process Equipment Design authored by Hesse and Rushton (1975), has been in used as the course textbook. In the succeeding paragraphs, calculation methods, conditions and data were reproduced from the said textbook.

O.4.1. Shell Design

Although in general shell thickness of 3/8 in is used for shell diameter between 12 and 24” (Peters and Timmerhaus, (1991), shell thickness of welded pressured vessel may be calculated using the given equation (Hesse and Rushton (1975).

(O – 38)

where tp = shell thickness (inch)P = Max allowable working pressure (psi)

O - 18


D = Inside diameter (inch)S = Max allowable tensile stress (psi) (Table O - 9)e = Efficiency of welded joint (Table O - 10)C = Corrosion allowance

The above equation is applicable as long as the following conditions are met:

1. tp < 0.10D2. tp > tmin

where

(O – 39)

O - 19


Table O – 9. Materials and Allowable Working Stresses for Unfired Pressure Vessels, Adapted from ASME-UPV Code.

SpecifiedASME Minimum Allowable Unit Tensile Stress, Thousands psiCode Tensile at Various Temperatures, °F

Spec. Material Data Strength - 20No. and Description Grad

e1000 psi to

650700 750 800 850 900 950 1000

S-2 Steel plates - flange and A 45 9.0 8.8 8.4 6.9 5.7 4.4 2.6firebox quality B 50 10.0 9.6 9.0 7.5 6.0 4.4 2.6

S-1 Carbon steel for boilers 11.0 10.4 9.5 8.0 6.3 4.4 2.5Carbon-silicon steel, A 55 11.0 10.4 9.5 8.5 7.2 5.6 3.8 2.0

S-42 ordinary strength range B 60 12.0 11.4 10.4 9.1 7.4 5.6 3.8 2.0S-44 Molybdenum steel A 13.0 13.0 13.0 12.5 11.5 10.0 8.0 5.0S-43 Low-carbon nickel steel AS-55 Carbon-silicon steel, high 65

strength range, 4-1/2” A 13.0 12.3 11.1 9.4 7.6 5.6 3.8 2.0plates and under

S-44 B 14.0 14.0 14.0 13.5 12.0 10.2 8.0 5.0S-43 B 70 14.0 13.3 11.9 10.0 7.8 5.6 3.8 2.0S-55 B 14.0 13.3 11.9 10.0 7.8 5.6 3.8 2.0S-44 C 15.0 15.0 15.0 14.4 12.7 10.4 8.0 5.0S-43 C 75S-28 Chrome-manganese-

siliconA 15.0 14.1 12.4 10.1 7.8 5.6 3.8 2.0

alloy steel B 85

Adapted from Hesse, H.E. and J.H. Rushton, Process Equipment Design

Design stress, S maybe estimated using the given equation:

S = Su x Fm x Fs x Fr x Fa (O – 40)

where Su = Minimum Specified Tensile StrengthFm = Material Factor

Fm = 1 for Grade A materialFm = 0.97 for Grade B material Fm = 0.92 for Grade C material

Fs = Temperature Factor (Use Table O - 11)Fr = Stress Relief (SR) Factor

Fr = 1.06 When SR is appliedFa = Radiographing Factor

Fa = 1.12 when Radiographing is applied and subsequent repair of defects

O - 20


Note: Both Stress Relief and Radiographing factors are equal to unity when not applied on welded joints.

Arc gas welding may induce internal strain and stress on welded joints. In this case, stress relieving such as by annealing or hammering may be employed to release localized stresses. A 6% increase in the allowable design stress is allowed in some cases.

Radiographing, on the other hand, is an application of X-ray on welded joints to examine defective fusion and other defects that may affect the integrity of the pressure vessel. If subsequent repair of a detected defect is done, a 12% increase in the allowable design stress may also be allowed.

Stress relieving is mandatory for:

1. tp > 1¼”

2. (For thinner plates)

where D has a minimum value of 20 inches3. ASTM A – 150 4. ASTM A – 149 (under certain conditions)

Whereas, Radiographing is mandatory for

1. ASTM A – 1502. ASTM A – 149 (under certain conditions)3. Lethal gases application4. Nuclear Reactor applications

Table O – 10. Types of Welded Joint and Corresponding Efficiencies.EFFICIENCY CRITERIA

LAP WELD (For circumferential Joint)

Single Lap Single Lap with plug weld Double Lap

BUTT WELD (For circumferential and longitudinal joints)

Single Butt Single Butt with Back-up Strip Double Butt Double Butt with reinforce at center

55%65%65%

70%80%80%90%

tp < ⅝”tp < ⅝”tp > ⅝”

tp < ⅝”tp < 1¼”tp > 1¼”tp > 1¼”

O - 21


Table O – 11. Material Factor

Metal Temperature, Plate and Forged°F Steel, % Cast Steel, %

Up to 650 25.0 16.7700 23.7 16.4750 21.0 14.7800 18.0 12.9850 15.0 11.1900 12.0 9.3950 9.0 7.51000 6.2 5.7

Adapted from Hesse, H.E. and J.H. Rushton, Process Equipment Design (1975)

In the recent American Society of Mechanical Engineers (ASME) Code (VIII-I), the working equation for the determination of shell thickness of cylinder subjected to internal pressure based on inside diameter is given as:

(O – 41)

where tp = shell thickness required (inch) [m]P = Internal pressure (psi) [kN/m2]R = Inside Radius (inch) [m]S = Allowable stress (psi) [kN/m2]E = Joint efficiency factor (Table O - 12)C = Corrosion allowance (inch) [m]

Provided that

1. tp less than or equal to and

2. Pressure is less than or equal to 0.385 SE (Jawad and Farr, 1988)

Corrosion allowances vary with fluid characteristics handled. For corrosive fluids, an allowance of ¼ inch and for non-corrosive, an allowance of ⅛ to inch are used (Backhurst and Harker, 1973).

O - 22


Figure O – 3. Welded Joint Categories.

Table O – 12. Maximum Allowable Joint Efficiencies1 for Arc and Gas Welded Joints.

Type No.

Joint Description Limitations Joint Category

Degree of Radiographic Examination

a b cFull Spot None

(1)

Butt joints as attained by double-welding or by other means which will obtain the same quality of deposited weld metal on the inside and outside weld surfaces to agree with the requirements of UW-35; welds using metal backing strips which remain in place are excluded.

None A, B, C & D

1.0 0.85 0.70

(2)Single welded butt joint with backing strip other than those included in (1)

(a) None except as shown in (b) below

A, B, C & D

0,90 0.80 0.65

(b) Circumferential butt joints with one plate offset, see UW-13(c) and Fig. UW-13.1 (k).

A, B & C 0.90 0.80 0.65

(3) Single-welded butt joint without use of backing strip

Circumferential butt joints only. Not over 5/8in. thick and not over 24in outside diameter

A, B & C NA NA 0.60

4)Double full fillet lap

joint longitudinal joints not over 3/8in. thick

A NA NA 0.55

O - 23


Double full fillet lap joint

circumferential joints not over 5/8in. thick

B & C NA NA 0.55

(5)

Single full fillet lap joints with plug welds confirming to UW-17

Single full fillet lap joints with plug welds confirming to UW-17

(a) Circumferential joints2 for attachment of heads not over 24in. outside diameter to shells not over 1/2in. thick.

B NA NA 0.50

(b) Circumferential joint for the attachment to shells of jackets not over 5/8in. in nominal thickness where the distance from the center of the plug weld to the edge of the plate is not less than 1-1/2 times the diameter of the hole for the plug.

C NA NA 0.50

(6) Single full fillet lap joints without plug

welds

(a) For the attachment of heads convex to pressure to shells not over 5/8in. required thickness. only with use of fillet weld on inside of shells, or

(b) For attachment of heads having pressure on either side. To shells not over 24in. inside diameter and not over 1/4in. required thickness with fillet weld on outside of head flange only.

A & B NA NA 0.50

1 E = 1.0 for butt joints in compression.2 joints attaching hemispherical heads to shells are excluded.Adapted from Jawad, Maan H., and James R. Farr, 1988, Structural Analysis and Design of Process Equipment, 2ed. John Wiley and Sons 1988.

Alternative ASME equation based on outside diameter of a cylindrical shell is given as:

(O – 42)

Shell Wall thickness for vacuum vessels may be calculated (Kalis, 1986) with this equation

O - 24


(O – 43)

where Pc = Collapsing pressure (psi)Te = Thickness to withstand external pressure (inch)Do = Outside diameter (inch)Em = Material’s modulus of elasticity

Te must be high enough so that Pc is five times greater than the difference between atmospheric pressure and design vacuum pressure

For Spherical Shell ,ASME code as cited by Kohan (1987) provide for equation to calculate the shell thickness:

(O – 44)

where P = Max allowable internal working pressure (psi)R = Inside Radius (inch)tp = Minimum required thickness (inch)E = Lowest joint efficiencyS = Max allowable stress (psi)

O.5. HEAT EXCHANGER

Figure O – 4. Shell-and- Tube Heat

Exchanger.O.5.1. Layout and Pitch Arrangement

O - 25


Tubes are usually arranged in a triangular or square pitch arrangement. Pitch is the center-to-center distance between tubes. Rotated square pitch, a variation of square pitch is the third commonly used tube arrangement as presented in Figure O – 5. While a triangular pitch arrangement offers more heat transfer area per unit volume of a heat exchanger, the square pitch arrangement offers ease in cleaning and maintenance operations. A minimum of 1.25 pitch to diameter ratio and/ or a minimum webb thickness between tubes of approximately 3.2 mm could ensure sufficient strength for tube rolling. Whereas a 6.4 mm clearance is suggested for mechanical cleaning requirement (Hewitt, et. al., 1994). In most design, the pitch to diameter ratio range from 1.25 to 1.5 (Peters et. al, 2004).

Figure O - 5. Tube Layout Patterns: (a) Square Pitch; (b) Triangular Pitch; (c) Square Pitch Rotated; (d) Triangular Pitch with Cleaning Lanes.

Tube layout normally follows symmetrical arrangement having the largest number of tubes at the center. With an appropriate pitch to diameter ratio and optimum pipe diameter chosen, known total heat transfer area, would lead to the shell diameter specification. Minimum shell diameter is calculated by:

Shell DiameterMin = Nc Do + (Nc + 1) C (O – 45)

where Nc = Number of tubes at the CenterC = Clearance

Clearance = Pitch – Diameter (O – 46)

O - 26


Table O – 16 . ASME Allowable Stresses for Some Alloyed Steel Material.

Spec. no.Nominal

composition P no.Group

no.Product

form Grade

Specified min. yield,

ksi

Specified min.

tensile, ksi

-20 to

100

Maximum allowable stress, ksi, for metal temp., °F,not exceeding:

200 300 400 500 600 700 750 800 850 900 950 1000 1050 1100 1150 1200SA-240 12Cr-1Al 7 1 Plate 405 25.0

60.0 15.0 14.3 13.8 13.3 12.9 12.4 12.1 11.7 11.1 10.4 9.7 8.4 4.0SA-268 12Cr-1Al 7 1 Smls. Tb. TP405 30.0SA-479 12Cr-1Al 7 1 Bar 405 25.0

SA-240 13Cr 7 1 Plate 410S30.0 60.0 15.0 14.3 13.8 13.3 12.9 12.4 12.1 11.7 11.1 10.4 9.7 8.4 6.4 4.4 2.9 1.8 1.0SA-268 13Cr 6 1 Smls. Tb. TP410

SA-268 12Cr-1Al 7 1 Wld. Tb. TP405 30.0 60.0 12.8 12.2 11.8 11.3 10.9 10.6 10.3 9.9 9.4 8.8 8.2 7.1 3.4

SA-268 13Cr 6 1 Wld. Tb. TP410 30.0 60.0 12.8 12.2 11.8 11.3 10.9 10.6 10.3 9.9 9.4 8.8 8.2 7.1 5.5 3.7 2.4 1.5 0.8SA-268 15Cr 6 2 Wld. Tb. TP429 35.0 60.0 12.7 12.1 11.7 11.3 10.9 10.5 10.2SA-268 17Cr 7 2 Wld. Tb. TP430 35.0 60.0 12.8 12.2 11.8 11.3 10.9 10.6 10.3 9.9 9.4 8.8 8.2 7.2 5.5 3.8 2.7 2.0 1.5

SA-268 11Cr-Ti 7 1 Wld. Tb. TP409 30.0 60.0 12.8 12.2 11.8 11.3 10.9 10.5 10.2 9.9 9.4SA-268 11Cr-Ti 7 1 Smls. Tb. TP409 30.0 60.0 15.0 14.3 13.8 13.3 12.9 12.4 12.1 11.7 11.1SA-268 18Cr-Ti 7 2 Wld. Tb. TPXM-8 30.0 60.0 12.8 12.2 11.8 11.3 10.9 10.6 10.3 9.9 9.4SA-268 18Cr-Ti 7 2 Smls. Tb. TPXM-8 30.0 60.0 15.0 14.3 13.8 13.3 12.9 12.4

SA-240 18Cr-Mo 7 2 Plate 18Cr-Mo 45.0 60.0 15.0 14.3 13.8 13.3 12.8 12.4SA-268 18Cr-Mo 7 2 Wld. Tb. 18Cr-Mo 45.0 60.0 12.8 12.2 11.8 11.3 10.9 10.5SA-268 18Cr-Mo 7 2 Smls. Tb. 18Cr-Mo 45.0 60.0 15.0 14.3 13.8 13.3 12.8 12.4

SA-240 13Cr 6 1 Plate 410 30.0 65.0 16.3 15.5 15.0 14.4 13.9 13.5 13.1 12.7 12.0 11.3 10.5 8.8 6.4 4.4 2.9 1.8 1.0SA-240 15Cr 6 1 Plate 429

30.0 65.0 16.3 15.5 15.0 14.4 13.9 13.5 13.1 12.7 12.0 11.3 10.5 9.2 6.5 4.5 3.2 2.4 1.8SA-240 17Cr 7 2 Plate 430SA-479 13Cr 6 1 Bar 410

40.0 70.0 16.2 15.4 14.9 14.4 13.9 13.4 13.1 12.6 12.0 11.2 10.4 8.8 6.4SA-182 13Cr 6 1 Forg. F6aCl.1SA-182 13Cr 6 3 Forg. Cl.F6aCl.2 55.0 85.0 21.3 20.3 19.6 18.9 18.2 17.6 17.1 16.5 15.7 14.4 12.3 8.8 6.4 4.4 2.9 1.8 1.0SA-217 13Cr 6 3 Cast. CA15 65.0 90.0 22.5 21.5 20.7 20.0 19.3 18.7 18.1 17.5 16.7 14.9 11.0 7.6 5.0 3.3 2.2 1.5 1.0SA-193 13Cr … … Bolt. B6(410) 85.0 110.0 21.2 21.2 21.2 21.2 21.2 21.2 21.2 21.2 19.5 15.6 12.0SA-268 15Cr 6 2 Smls. Tb. TP429

35.0 60.0 15.0 14.3 13.8 13.3 12.9 12.4 12.1 11.7 11.1 10.4 9.7 8.5 6.5 4.5 3.2 2.4 1.8SA-268 17Cr 7 2 Smls. Tb. TP430SA-479 17Cr 7 2 Bar TP430

40.0 70.0 17.5 16.6 16.1 15.5 15.0 14.5 14.1 13.6 12.9 12.1 11.0 9.2 6.5SA-479 18Cr-Ti 7 2 Bar TPXM-8

SA-268 26Cr-4Ni-Mo 10E 5 Wld. Tb. TP329 70.0 90.0 19.1 19.1 18.4 18.0 18.0SA-268 26Cr-4Ni-Mo 10E 5 Smls. Tb. TP329 70.0 90.0 22.5 22.5 21.6 21.2 21.2SA-240 26Cr-4Ni-Mo 10E 5 Plate TP329 70.0 90.0 22.5 21.9 20.5 19.8 19.8

SA-268 27Cr 10E 5 Smls. Tb. TP446 40.0 70.0 17.5 16.6 16.1 15.6 15.0 14.5

SA-412 17Cr-4Ni-6Mn 8 1 Plate 201 45.0 95.0 23.8 20.8 19.1

SA-182 18Cr-8Ni 8 1 Forg. F304L 25.0 65.0 15.5 15.4 14.2 13.6 13.4 13.3 13.1 13.0 12.9SA-240 18Cr-8Ni 8 1 Plate 304L

25.0 70.0 15.7 15.7 15.3 14.7 14.4 14.0 13.5 13.3 13.0SA-213 18Cr-8Ni 8 1 Smls. Tb. TP304LSA-312 18Cr-8Ni 8 1 Smls. Tb. TP304LSA-479 18Cr-8Ni 8 1 Bar 304L

Adapted: ASME code, Section VIII, Division 1.

O - 27


where

Gt = Mass velocity in the tubes (O – 52)

(O – 53)

N = Total number of tubesn = Number of tube side passes

a1 = Cross-sectional area of flow per tube

Dt = Tube diameter

= Fluid velocity

f = Friction Factor

(O – 54)

(O – 55)

Applicable for NRe >1000

B. Shell side

(O – 56)

(O – 57)

where Gs = Mass velocity in the shell

Ds = Shell inner diameter [ft] [m]S = Specific gravity of the fluid

O - 28


= (O – 58)

L = Tube length [ft] [m]B = Baffle spacing [ft] [m]

f = Friction factor

(O – 59)

(O – 60)

Applicable for NRe >500

Peters and Timmerhaus, (1991) provide an alternative equation for pressure drop across the tube and shell as reproduced in the following paragraph:

For tube – side

(O – 61)

where subscript i refers to inside of tube at bulk temperaturefi = Fanning friction factor for isothermal flow based on conditions at the

arithmetic-average temperature of the fluidnp = Number of tube passesgc = Conversion factor in Newton’s law of motion,

(O – 62)

ΦI = Correction factor for non-isothermal flow

(O – 63)

O - 29


when Di/μ is less than 2100 and

(O – 64)

when is greater than 2100;

i = Viscosity at arithmetic – average (bulk) temperature of fluidw = Viscosity of fluid at average temperature of the inside tube wall surfaceBi = Correction factor to account for friction due to sudden contraction, expansion

and reversal of flow direction

(O – 65)

For flow across tubes, the following equation can be used to approximate pressure drop due to friction:

(O – 66)

where subscript o refers to outside of tube at bulk temperature

f1 = Special friction for shell-side flow

(O – 67)

(O – 68)

(O – 69)

(O – 70)

(O – 71)

O - 30


Nr = Number of rows of tubes across which shell fluid flowsBo = Correction factor to account for friction due to reversal in directional flow

recrossing of tubes, and variation in cross section1 when the flow is across unbaffled tubes orNumber of tubes crosses as a rough approximation

Variation of Kern method and other estimations by Bell-Delaware method and Wills and Johnson method are discussed in Process Heat Transfer (Hewitt, et. al., 1994).

O.5.4. Heat Exchanger Temperature Limits

The most common heat exchanger medium used for cooling is water. Aside from its abundance and cost, water exhibits relatively high heat capacity. In the design of heat exchanger, it is obvious that either large quantity of cooling will be used or greater water temperature change should be anticipated to come up with smaller heat exchanger. Large quantity of cooling water would result to higher water velocity. This high velocity will reduce fouling but increases water and pumping costs. On the other hand, large water temperature increase will require less water and pumping costs. However, at high temperatures, water exerts considerable corrosive action on steel, particularly if water contains dissolved oxygen (Peters et. al, 2004). Furthermore at high water temperature scaling tends to increase (Backhurst and Harker, 1973). To minimize scale formation, water temperature should not be more than 120ºF (Backhurst and Harker, 1973; Peters et. al, 2004). To protect against fouling and corrosion, water temperature (outlet) should not be heated above 158F (Baasel, 1974). Again a good compromise has to be set between large quantity of cooling water and greater water temperature change.

For the cooling water, on an open circulation systems such as cooling towers and spray ponds, the temperature of the cooled water is 8-13ºF above the wet bulb temperature (Baasel, 1974). However since oxygen is picked-up in every pass, treatment of water is necessary if corrosion and growth of microorganism is to be controlled (Peters et. al, 2004).

When using cooling water to cool or condense a process stream, assume a water inlet temperature of 90oF (from a cooling tower) and a maximum water outlet temperature of 120oF (Seider et al, 2004).

As to the temperature difference, the rule of thumb is that the greatest temperature difference in an exchanger should be at least 36F and the minimum temperature difference should be at least 10oF hot (Lord et. al., 1970).

O - 31


O.6. CSTR DESIGN

O.6.1. CSTR Sizing

CSTR sizing is dictated by residence time requirement. The longer the residence time, the bigger the reactor volume at constant volumetric flow rate. This is expressed below:

τ = V / Qυ (O – 72)

where: τ = Space time or Residence time, sec [hr]V = Volume of Reactor, m³ [ft3]Qυ = Volumetric flowrate, m³/sec [ft3/s]

Overall chemical kinetics which includes, chemical specie, amount of specie, reaction temperature, presence of catalyst, agitation etc determines the degree of residence time as shown in Table O – 17.

Table O – 17. Residence Time and/or Space Velocities in Industrial Chemical Reactors.

Product(raw materials)

Reactor Phase

(CSTR)Catalyst T, °C

P, atm

Residence Time or Space

VelocityAlkylate (i-C4, butanes) L H2SO4 5-10 2-3 5-40 minAlkylate (i-C4, butanes) L HF 25-38 8-11 5-25 minButadiene sulfone (butadiene, SO2)

L t-Butyl catechol 34 12 0.2 LHSV

Caprolactam (cyclohexane oxime)

L Polyphosphoric acid

80-110

1 0.25-2 h

Chloral (Cl2, acetaldehyde) LG None 20-90 1 140 hCumene hydroperoxide (cumene, air)

L Metal porphyrins 95-120

2-15 1-3 h

Cyclohexanone (cyclohexanol)

L N. A. 107 1 0.75 h

Dextrose (starch) L H2SO4 165 1 20 minDextrose (starch) L Enzyme 60 1 100 minDodecylbenzene (benzene, propylene tetramer)

L AlCl3 15-20 1 1-30 min

Ethyl acetate (ethanol, acetic acid)

L H2SO4 100 1 0.5-0.8 LHSV

Ethylene, propylene chlorohydrins (Cl2, H2O)

LG None 30-40 3-10 0.5-5 min

Glycerol (allyl alcohol, L H2WO4 40-60 1 3 h

O - 32


H2O2)o-Methyl benzoic acid (xylene, air)

L None 160 14 0.32 h3.1 LHSV

Nitrobenzene (benzene, HNO3)

L H2SO4 45-95 1 3-40 min

Phenol (cumene hydroperoxide)

L SO2 45-65 2-3 15 min

t-Butyl methacrylate (methacrylic acid, i-butene)

L H2SO4 25 3 0.3 LHSV

Aldehydes (diisobutene, CO)

LG Co Carbonyl 150 200 1.7 h

LHSV – Space velocity (hourly) – liquid N. A. – Not Available Adapted from Table 23 - 1 Section 23 - 7 Perry’s Chemical Engineer’s Handbook 7th ed.

O.6.2. Standard Stirred Tank Configuration

Trambouze et. al., (1988), provide for the standard stirred tank configuration for a six flat blade turbine type agitation system. As shown below, the following are standard configuration:

Figure O – 7. Dimensions for CSTR Design.

O - 33

Dd

DI

m

I

b

ZI

ZL

W

HT

DT


(O – 75) (O – 76)

(O – 77) (O – 78)

(O – 79) (O – 80)

(O – 81) (O- 82)

where: = Static liquid depth

= Tank diameter

= Impeller diameter

= Impeller distance from tank bottom = Baffle width

= Impeller disc diameter = Impeller blade length

= Impeller blade width= Tank height= Baffle tip distance from tank bottom

O.6.3. Mixing Time

To estimate the mixing time, Norwood and Metzner correlation provides for the equation applicable for six flat blade turbine:

(O – 83)

where: = Mixing time= Impeller revolutions per unit time= Impeller diameter

= Tank diameter

= Static liquid depth

= Froude Number

O - 34


(O – 84)

where g = acceleration due to gravity

O.6.4. Impeller Selection

Agitation is designed to increase fluid turbulence, and is often employed in the following:1. homogenization of a fluid phase2. increased heat transfer between a solid surface and a fluid phase3. creation of interfacial area between two immiscible fluid phases.4. maintenance of a divided solid in suspension in a fluid phase

Agitation as used in the process industries is the production of irregular disturbances or turbulent motion within a fluid by means of mechanical devices acting on that fluid (Brown, 1950). Most of the fluids handled in the process industry are low viscosity Newtonian fluids.

Several references classify impellers according to their form, functions and uses in the mixing operations (Brown, 1950; Foust et. al, 1980 and McCabe, 2001) as shown in Figure O – 8. In the selection of appropriate impeller type Figure O – 9 may be used.

Figure O – 8. Types of Impeller

Source: Doran, Pauline M. 1995 Bioprocess Engineering Principles.

A graphical method of impeller selection is presented on Figure 0 - 9.Figure O – 9. Viscosity Ranges for Different Impellers

O - 35

Anchor Propeller 6 flat blade disc-turbine

Paddle Gate anchor Helical screw

Impeller Type

Vis

cosi

ty (

cen

tip

ois

e)

103

104

105

106

107

102

10

1

An

chor

s

Pro

pel

lers

Fla

t-b

lad

e tu

rbin

es

Pad

dle

s

Gat

e an

chor

s

Hel

ical

Scr

ews

Hel

ical

Rib

bon

s


Source: F.A Holland and F.S. Chapman, 1966, Liquid Mixing and Processing in Stirred Tanks as cited by Doran, Pauline M. 1995. Bioprocess Engineering Principles)

O.6.5. Baffles

Baffles are flat vertical strips set radially along the tank inner wall. They are mounted inside the tank to produce higher mixing and horizontal liquid surface (Perry and Green, 1997). In the absence of baffles in a stirred tank, vortex are formed because of the centrifugal force acting on the liquid and could reach deep to the impeller which is undesirable (Mc Cabe, 2001). Due to the motion of the impeller in the fluid and the resultant movement of the liquid past the baffles and wall, the skin friction and the drop form have to be considered in relation to the speed of rotation and design of blade and tank.

Figure O – 10 provides for baffle inclination and attachment selection guide.

Figure O – 10. Baffle Arrangements (a) Baffles are attached to the wall for low-viscosity liquids. (b) Baffles set away from the wall for moderate-viscosity liquids. (c) Baffles set away from the wall and at an angle for high-viscosity liquids.

O - 36


Source: F.A Holland and F.S. Chapman, 1966, Liquid Mixing and Processing in Stirred Tanks as cited by Doran, Pauline M. 1995. Bioprocess Engineering Principles)

O.6.6. Power Dissipation

Power dissipated by the agitator maybe computed by:

(O – 85)

where Pa = Power dissipated by an agitatorN = RPM of the impellerρ = Density of the mixtureDI = Impeller diameterNP = Power number

An estimation of typical horsepower for agitators is given below (Parker, 1964; Schlegel, 1972): This maybe used to approximate power requirement due to mixing of CSTR.

Fluid Approximate horsepower

O - 37


Blending vegetable oil 1.0 hp per 100,000 lbBlending gasoline 0.019 hp per m3

Clay dispersion 10 – 12 hp per 1,000 galFermentation (pharmaceutical) 3 – 10 hp per 1,000 galSuspension polymerization 6 – 7 hp per 1,000 galEmulsion polymerization 3 – 10 hp per 1,000 galSolution polymerization 15 – 40 hp per 1,000 gal

Radius of Action of an Agitator

Radius of action of an agitator should be checked after reactor, blade and baffle sizes have been calculated to ensure there is enough intensity of mixing inside the reactor, as this will affect reaction conversion. Radius of Action may be calculated as:

(O – 86)

Horizontal radius of action and vertical radius of action are 50% and 20% respectively of the computed radius of action.

where = Power, watts = Viscosity, Pa . s = Radius of action, m

= Half major axis ellipsoidal = Half minor axis ellipsoidal

To ensure high degree of agitation a linear speed at blade tip should be greater than 4. Where tip speed is given by:

Vp = πNDI (m/s) (O – 87)

Another indicator of high degree of agitation is Power dissipated per unit volume of fluid

which should have at least 1,500 value.

O - 38


Below is the summary of degree of agitation against tip blade speed and Power per unit volume (Trambouze et. al, 1988):

Degree of Agitation Tip Speed

Low 3.25 750Medium 3.25 to 4 750 to 1500High 4 up 1500 up

For an initial condition, a 50% on blade tip speed of 4 and Power per unit volume of

1500 could be a good choice. On the other hand, a good compromise should be

reached, so that just enough mixing is provided for certain required residence time for power requirement to be justifiable. An acceptable criteria used is:

tm τ

O - 39

< 0.1


allowable stress, O - 9, 30Annular Plates, O - 17API, O - 12API 650 tanks, O - 13ASME, O - 6, 8, 9, 21, 24, 28, 30AWWA, O - 13cast iron pipe, O - 8clean fluids, O - 1Corrosive Fluids, O - 2CSTR design, O - 35CSTR design, baffles, O - 27, 41CSTR design, sizing, O - 35CSTR design, impeller selection, O - 38CSTR design, mixing time, O - 38CSTR design, power dissipation, O - 42CSTR design, configuration, O - 37CSTR design, types of reactor, O - 35erosive Fluids, O - 2equipment design, O - 1heat exchanger, O - 25heat exchanger, baffle and tube sheet, O - 27heat exchanger, temperature limits, O - 34heat exchanger, pressure drop across the, O - 31heat exchanger, layout and pitch arrangement, O - 25joint efficiencies, O - 9, 14, 20, 21, 22, 28, 30other pipe materials of construction, O - 10pipe and tube sizing, O - 1pipe and tube thickness, O - 6pipe diameter, economic, O - 2pipe diameter, typical, O - 1pitch. See heat exchanger layoutstorage tank design, O - 11storage tank design, shell thickness, O - 12storage tank design, sizing, O - 11TEMA, O - 28, 29welded joint, O - 20, 22welded pressured vessel (unfired), O - 18welded steel, O - 12, 13

O - 40

Process Equipment Design (Pinoy Edition)

Documents

pipe x

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choice of pipe diameter

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