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Note: The source of the technical material in this volume is the Professional

Engineering Development Program (PEDP) of Engineering Services.

Warning: The material contained in this document was developed for Saudi

Aramco and is intended for the exclusive use of Saudi Aramco’semployees. Any material contained in this document which is not already

in the public domain may not be copied, reproduced, sold, given, or

disclosed to third parties, or otherwise used in whole, or in part, without

the written permission of the Vice President, Engineering Services, Saudi

Aramco.

Chapter : Process For additional information on this subject, contact

File Reference: CHE10110 R. A. Al-Husseini on 874-2792

Engineering Encyclopedia Saudi Aramco DeskTop Standards

Other Heat Transfer Equipment

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Engineering Encyclopedia Process


Saudi Aramco DeskTop Standards

CONTENTS PAGES

PROCESSING STEPS IN REFRIGERATION SYSTEMS ..................................1

Determining Refrigeration Systems Requirements ....................................5

Calculating Percent of Design Duty ...........................................................9

Air-Cooled Exchangers ..............................................................................9

Fan Types .................................................................................................10

Air Control ...............................................................................................12

Ambient Air Effects..................................................................................12

CALCULATING COOLER DESIGN REQUIREMENTS..................................15

Cooler Configuration/Materials................................................................15

Tube Fin Types.........................................................................................15

Cooler Design Method .............................................................................19

Calculating Plot Area ...............................................................................30

OTHER TYPES OF HEAT EXCHANGERS ......................................................32

General .....................................................................................................32

Plate/Frame...............................................................................................34

Hairpin......................................................................................................37

Box Coolers..............................................................................................39

Enhanced Heat Transfer ...........................................................................41

CALCULATING FURNACE EFFICIENCY ......................................................45

General .....................................................................................................45

Direct Fired...............................................................................................46

Forced Draft Furnaces ..............................................................................52

Combustion Air Preheaters.......................................................................52

Efficiency .................................................................................................52

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Saudi Aramco DeskTop Standards

Part 2 - Calculate Furnace Efficiency.......................................................54

Routine Furnace Startup and Operations..................................................58

Startup...........................................................................................58

Optimum Excess Air Levels .........................................................61

Monitoring Devices and Techniques........................................................65

Controls/Safety Devices/Burners..............................................................65

HEATER ALARM/SHUTDOWN SYSTEM DESCRIPTION............................67

Monitoring Tube Metal Temperature .......................................................71

NOMENCLATURE.............................. ...............................................................73

KEY FORMULAS...............................................................................................75

WORK AID 1 - PROCEDURES FOR CALCULATING PERCENT OF

DESIGN DUTY.........................................................................77

WORK AID 2 - PROCEDURES FOR CALCULATING EXTENDED

SURFACE AND FACE AREA REQUIREMENTS..................79

WORK AID 3 - PROCEDURES FOR CALCULATING FURNACE

EFFICIENCY.............................................................................82

GLOSSARY ........................................................................................................84

REFERENCES.....................................................................................................87

APPENDICES .....................................................................................................88

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Saudi Aramco DeskTop Standards 1

PROCESSING STEPS IN REFRIGERATION SYSTEMS

Refrigeration systems are used to lower the operating temperature of certain processing

schemes to temperatures that cannot be obtained via air and/or water cooling. The lower

process operating temperatures are usually required because of:

• The process fluid's thermodynamic properties.

• Fouling tendencies at higher temperatures.

• The economic attractiveness of operating the process at a lower temperature.

• The system refrigerant is usually selected based on:

• The refrigerator temperatures needed.

• The type of refrigerants available from the process plant.

• A need to not contaminate the process plant fluid if equipment were to leak.

Some common refrigerants and their properties are listed below.

ASHRAE

Refrigerant

Number

Chemical

Name

Chemical

Formula

Molecula

r Weight

Normal

Boiling

Point °F

@ 14.696

psia

Critica

l

Temp.

°F

Critical

Press-

ure psia

Freezing

Point °F

@ 14.696

psia

Liquid

Viscosity

Centipoise

Liquid

Thermal

Conductivity

Btu

(hr ft2 °F) ft

Specific

Heat

Ratio k =

C p/Cv

Toxicity

UL

Group

Classifi-

cation

11

Trichloro-

fluoromethane CC13F 137.4 74.8 388.4 640.0 -168

0.421@NBT

0.395@86°F

0.0506@NBT

0.0498@86°F 1.13 5

114

Dichlorotetra-

fluoroethane CC1F2OC1F

2

170.0 38.4 294.3 474.0 -137

0.44@NBT

0.32@86°F

0.0405@NBT

0.0366@86°F 1.09 6

12

Dichlorodifiuoro

Methane CC12F2 120.9 -21.6 233.6 597.0 -252

0.358@NBT

0.206@86°F

0.0518@NBT

0.0392@86°F 1.14 6

22

Chlorodifiuoro

Methane CHC1F2 86.5 -41.4 204.8 716.0 -256

0.33@NBT

0.192@86°F

0.0695@NBT

0.0495@86°F 1.18 5a

600 N-Butane C4H10 58.1 31.1 305.6 550.7 -217

0.213@NBT

0.159@86°F

0.0663@NBT

0.061@86°F 1.09 5b

290 Propane C3H8 44.1 -43.7 206.0 616.3 -305

0.21@NBT

0.101@86°F

0.076@NBT

0.056@86°F 1.14 5b

1270 Propylene C3

H6

42.1 -53.9 197.1 667.2 -301

0.15@NBT

0.089@86°F

0.082@NBT

0.057@86°F 1.15 5b

170 Ethane C2H6 30.1 -127.4 9.01 707.8 -297

0.168@NBT

0.039@86°F

0.082@NBT

0.048@86°F 1.19 5b

1150 Ethylene C2H4 28.1 -154.8 48.6 731.1 -272

0.17@NBT

0.07@86°F

0.111@NBT

0.031@86°F 1.24 5b

50 Methane CH4 16.0 -258.7 -116.7 667.8 -296 0.118@NBT 0.110@NBT 1.305 5b

717 Ammonia NH3 17.0 -28.0 270.4 1636.0 -108

0.25@5°F

0.207@86°F

0.29@32°F

0.26@86°F 1.29 2

With permission from the Gas Processors Suppliers Association. Source: Engineering Data

Book.

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Table 1

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A very simplified process flow diagram showing how a refrigeration system works is shown

in Figure 1. The refrigerant is compressed to a pressure that permits condensing the

compressor discharge with water or air coolers. The condensed liquid is then fed to the

process cooler (chiller or evaporator). An expansion valve on the refrigerant feed to the

process cooler significantly lowers the pressure of the refrigerant. The depressured refrigerant

in the cooler boils at a low temperature, thereby cooling the process fluid on the tube side of

the cooler. Low-pressure vapors from the process cooler shell flow back to the compressor.

The refrigeration cycle just discussed (from the compressor through the water or air cooler,

expansion valve and process fluid cooler, back to the compressor) is also shown

thermodynamically on the pressure/enthalpy diagram in Figure 1. The A, B, C points on the

pressure/enthalpy diagram correspond to the letters on the process flow diagram in Figure 1.

Figure 2 is a pressure/enthalpy diagram that would be used by an engineer to review an

operating refrigeration system or design a new system.

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PROCESS FLOW DIAGRAM

Flow Diagram

PRESSURE ENTHALPY DIAGRAM

With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.

Figure 1

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Determining Refrigeration Systems Requirements

Engineers will often review pressure/enthalpy diagram when reviewing operating

refrigeration systems or designing new systems. The various calculations involved with the

four processing steps discussed earlier can also be determined using these diagrams. Figure 2

is an example of a diagram from which all four refrigeration processing steps, compression,

condensation, refrigerant flashing, and vaporization may be determined.

Figure 2, a Propane P-H diagram, is used for refrigerant systems calculations in which

propane suitable for refrigeration is manufactured during a process. The primary outcome of

these calculations is to determine the quantity of refrigerant needed to provide the appropriate

process cooling duty.

The enthalpy of the condensed compressor discharge can be read from Figure 2 at the

intersection of the bubble point curve and the known condensing temperature of water in thesystem.

After the appropriate compressor suction pressure is selected (this pressure will boil the

propane at a low enough temperature to ensure the desired temperature difference across the

tubes in a process cooler), the enthalpy of the vapor returning to the compressor may be read

at the intersection of the psia pressure line and the dew point line. The enthalpy of the liquid

refrigerant (in this case, propane) to the process cooler is read at the intersection of the psia

line and the bubble point curve. Note that the lb/hr of refrigerant vaporized in the process

cooler is determined using the following formula:

Process DutyRefrigerant Heat of Vaporization

After the lb/hr of refrigerant vaporized is determined, the lb/hr of refrigerant flashed across

the expansion valve must be calculated. The refrigerated flash is an isenthalpic reaction and

is calculated using the process cooler refrigerant rate and the enthalpies of the feed and

products to and from the expansion valve. It is important to remember that the total pounds of

refrigerant from the valve is equal to the total pounds to the valve.

The calculation for refrigerant flashed lb/hr can be summarized as:

(Enthalpy of Compressor Discharge Condensed Liquid, Btu/lb) (Compressor, Gas Rate, lb/hr)= (Vapor from Pressure Reduction Valve, lb/hr) (Vapor Enthalpy, Btu/lb) + (Liquid from

Pressure Reduction Valve, lb/hr) (Liquid Enthalpy, Btu/lb)

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PRESSURE/ENTHALPY DIAGRAM


Figure 2

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This calculation will yield the lb/hr of refrigerant needed to provide process duty at the

specified temperature. Thermodynamic tables for refrigerants are available and can be used

in place of the P-H diagram (Figure 2) for enthalpy data. When selecting the appropriate

refrigerant, the following factors should be considered:

• Atmospheric boiling point.

• Availability.

• Fire hazard possibilities.

• Toxicity.

For additional information, Figure 3 illustrates how a simple system can be expanded into a

complex one with multiple stages and refrigerants. Figure 4 shows one of several types of

curves available for use by engineers in the design and evaluation of refrigeration systems.

In summary, the four processing steps in a refrigeration system can be calculated using a P-Hdiagram.

CASCADE REFRIGERATION SYSTEM


Figure 3

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SINGLE-STAGE PROPANE REFRIGERATION SYSTEM

(use photostat)


Figure 4

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Calculating Percent of Design Duty

Calculating the percent of design duty a cooler can perform during normal or abnormal

operations involves a number of calculations. The next three sections, fan types, air control,

and ambient air effects will provide sufficient information needed for calculations of percent

of design duty.

Air-Cooled Exchangers

The extent to which air-cooled exchangers are used is one of the first major economic

evaluations made in connection with a process system design. Generally, air-cooled

exchangers are more attractive for the higher temperature heat removal services, and water-

cooled exchangers are more attractive for the lower temperature heat removal services. An

economic evaluation will establish the break point temperature for the change from air-cooled

to water-cooled exchangers. The availability of plot space for air coolers and the availabilityof makeup water for the cooling water system also significantly influence the air versus water

evaluation, particularly for unusual circumstances.

The design alternatives available for specifying an air-cooled exchanger will now be

discussed. Initially, an engineer should refer to Saudi Aramco documents AES-E-001 and

ADP-E-001 for the preferred air-cooled exchanger configurations used by Saudi Aramco.

Whenever possible, the final design should conform to these preferred configurations. See

Figure 5 for a typical air-cooled installation.

TYPICAL SIDE ELEVATIONS OF AIR COOLERS

Figure 5

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TYPICAL PLAN VIEWS OF AIR COOLERS

Figure 5 (Cont'd)

Fan Types

Air-cooled exchangers can be specified with induced draft or forced draft fans. If the fan is

located below the tube bundle and the fan discharges upward, the cooler has a forced draft

fan. If the fan is above the tube bundle and discharges upward, the cooler has an induced fan.

The advantages and disadvantages of each are listed below.

The advantages of induced draft fans are:

• Better distribution of air across the section.

• Less possibility of the hot effluent air recirculating around to the intake of the sections.

The hot air is discharged upward at approximately 2-1/2 times the velocity of intake, or

about 1500 ft/min.

• Less effect of sun, rain, and hail, since 60% of the face area of the sections is covered.

• Increased capacity in the event of fan failure, since the natural draft stack effect is much

greater with induced draft.

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• With permission from the Gas Processors Suppliers Association. Source: Engineering

Data Book.

The disadvantages of induced draft fans are:

• Higher horsepower, since the fan is located in the hot air. (Horsepower varies directly

with the absolute temperature.)

• Effluent air temperature should be limited to 200°F to prevent potential damage to fan

blades, bearings, V-belts, or other mechanical components in the hot air stream.

• The fan drive components are less accessible for maintenance, which may have to be done

in the hot air generated by natural convection.

• For inlet process fluids above 350°F, forced draft design should be used; otherwise, fanfailure could subject the fan blades and bearings to excessive temperatures.

The advantages of forced draft fans are:

• Slightly lower horsepower because the fan is in cold air.

• Better accessibility of mechanical components for maintenance.

• Easily adaptable for warm air recirculation for cold climates.

• Capable in cooling process fluids with higher inlet temperatures.

The disadvantages of forced draft fans are:

• Poor distribution of air over the section.

• Greatly increased possibility of hot air recirculation, due to low discharge velocity

• from the sections and absence of stack.

• Low natural draft capability on fan failure, due to small stack effect.

• Total exposure of tubes to sun, rain, and hail.

With permission from the Gas Processors Suppliers Association. Source: Engineering DataBook.

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Air Control

It is often attractive to regulate the air flow to the airfin cooler and/or condenser because

seasonal variations in air temperature can result in considerable overcooling of the processstream. Overcooling can cause exchanger plugging or other processing problems. The air

can be regulated by variable speed drives for the fan or variable pitch blades on the fans.

Louvers on the face of the tube bundle will also control the flow of air. Louvers will not

reduce fan motor electrical consumption, whereas the variable fan pitch and variable speed

drive will.

Since the air-cooled exchanger must be specified for the hottest time of the year (typically 94

to 96% of the maximum expected air temperature, +5°F to allow for air recirculation), for

most of the year there is an opportunity to reduce electrical costs with variable pitch blades or

a variable speed motor. Designing the air cooler at 94 to 96% of the maximum temperature

versus the maximum air temperature often reduces the cost of the air cooler by 50 to 60%.

Refer to Saudi Aramco AES-E-002 (4-14-3) for Saudi Aramco design air temperature.

Appendix A is a sample of the input to and output from a Saudi Aramco computer program

from an airfin exchanger.

Ambient Air Effects

Changes in air temperature significantly affect an operating air cooler's process cooling

capability and the horsepower requirement for the fans. Example Problem 1 shows that when

the air temperature is at the minimum design level, about half the airfin cooling surface onthis service could be shut down to compensate for the Æte increasing from 75 to 164°F.

Actually, the partial shutdown mode is accomplished by shutting down some of the airfin fans

on a large duty air cooler that has three or more fans during the winter, when air temperatures

are consistently below design. Shutting down fans or specifying variable pitch fan blades or

drivers for the fans can often result in 20 to 80% of rated horsepower savings, annually, due

to seasonal changes in the air temperature. The effect of changes in ambient air temperature

on cooling capacity is illustrated by Example Problem 1.

Example Problem 1

In order to calculate the percent of design duty, Æte must first be calculated. Use Part 1 of Work Aid 1 to calculate Æte for the design air temperature operation and the minimum air

temperature application, using FT factors based on curves in Figure 6. Assume that the air

cooler is cooling a product stream from 280°F to 170°F, the exchanger duty is 80 MBtu/hr,

the design air temperature is 120°F, the minimum air temperature is 38°F (night time in the

winter), and the air temperature rise across the cooler is 45°F when inlet air to the air cooler is

at the design value (120°F). Also assume that the air cooler has four tube rows.

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CORRECTION FACTORS FOR LIMITED CROSS FLOW EXCHANGERS

(use photostat)

Basis: X =T1 - T2

T1 - t1 Y =

t2 - t1T1 - T2

where: t1 = Ambient Air, °F

t2 = Average Hot Air, °F

T1 = Tubeside Inlet, °F

T2 = Tubeside Outlet, °F

Corrected LMTD = (FT) (tm) = Æte

Figure 6

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CALCULATING COOLER DESIGN REQUIREMENTS

Cooler Configuration/Materials

Air cooler configuration is greatly influenced by the type of exchanger that Saudi Aramcouses as a standard design (see ADP-E-001). The standard size is 1-in. O.D. tubes, and 1 1/2-

in. O.D. tubes are used for high viscosity fluids. The standard tube lengths are 12-, 20-, 24-,

30-, 32-, 36-, or 40-ft long. Standard fins are aluminum and have a height in the 3/8- to 5/8-

in. range (10 fins per linear inch of tube), depending on the quality of the fluid being cooled

(hi value). The higher the tubeside coefficient (hi) is, the more the optimum height of the fins

increases. Saudi Aramco standardizes on exchangers that have tube rows in the range of 3 to

8 deep. As the temperature difference between the process fluid inlet or air inlet temperatures

increases or the overall coefficient (Uo) decreases, the optimum number of tube rows

increases (see Figure 7).

Figure 7 should be used only for the exchanger configuration noted on the figure. This type

of exchanger is one of the standard Saudi Aramco configurations. The optimum number of

tube rows indicated in Figure 7, as well as the allowable tubeside pressure drop, influences

the number of tube passes selected. For tube pitch, Saudi Aramco has standardized on a 2

3/8-in. triangular spacing. This standard tube pitch should always be used, except for unusual

services. An unusual service would be a circumstance where the 2 3/8-in. triangular pitch, in

combination with the specified number of tube rows, face area, and air rate, results in a static

air pressure requirement too high for the fans. Air fan static head availability covers only a

narrow range. The maximum permissible value for static head is about 0.7 in. H2O with a

normal value being closer to 0.5 in. H2O. The acceptable static head range corresponds to a

face velocity range of 6.5 to 13 ft/s, depending on the temperature of the air and the number of tube rows in the exchanger. The preferred way of correcting an air pressure drop problem

would be to change the face area of the air-cooled exchanger or the number of tube rows

before changing the standardized tube pitch.

Air-cooled exchanger materials for the tubes and headers (parts in contact with the process

fluid) are specified in the same manner as shell and tube exchangers. A tubewall and header

wall thickness is selected so that at the end of the desired equipment life (say 10 years or 15

years) the corroded wall will still be thick enough to contain the process fluid pressure.

Tube Fin Types

Saudi Aramco has standardized on two types of air cooler tube fins:

• 5/8-in. high fins, 10 fins/inch of tube, extruded fin type.

• 5/8-in. high fins, 11 fins/inch of tube, embedded fin type.

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Both types are made of aluminum. Other types of fins available are wrapped fins that fit into

spiral grooves cut into the host tube or fins first wrapped around the tube. The fins may also

be serrated or plain. Some of the different types of tube fins are shown in Figure 8.

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EFFECT OF TEMPERATURE LEVEL AND OVERALL HEAT

TRANSFER RATE UPON OPTIMUM BUNDLE DEPTH

(use photostat)

Basis: 1-in. O.D. x 24-in. long steel tube with extruded aluminum fins and 2 3/8-in. triangular spacing.

where: T1 = Inlet temperature of fluid to be cooled, °F.

t1 = Inlet air temperature.

Uo = Overall heat transfer coefficient (related to bare tube outside diameter) Btu/hr ft2 °F.

Figure 7

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TYPES OF FINNED TUBES USED IN AIR-COOLED HEAT EXCHANGERS

Figure 8

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Cooler Design Method

Saudi Aramco Design Practice ADP-E-001 contains a simplified design calculation procedure

that be used to check vendors' proposals for air-cooled exchangers or review the performance

of an operating exchanger. The Saudi Aramco design practice procedure will be used for

Example Problem 2 below, which will show how to review a vendor-specified air-cooled

exchanger, checking the adequacy of the vendor-specified surface, face area, and fan

horsepower.

Example Problem 2

Oftentimes vendors will specify air-cooled exchangers. In order to determine if the vendor

specified exchanger meets Saudi Aramco design expectations, a series of simplified design

calculations can be used. These calculations can also be used to review the performance of

operating exchangers. The following example problem will demonstrate how to review avendor-specified air-cooled exchanger, checking the adequacy of the vendor-specified

surface, face area, and fan horsepower.

It is necessary to check the following vendor-proposed air-cooled exchanger. Calculate the

air temperature rise, air rate, surface (extended) requirement, face area, number of fans, and

horsepower for each fan to confirm the vendor's design. Refer to Work Aid 2 for procedures

to calculate extended surface and face area requirements.

Assume the following:

Exchanger duty Q = 31 Mbtu/hr.

Process fluid temperature in/out = 250/150°F.

Process fluid flow rate Wt is 550,000 lb/hr of hydrocarbon with a specific heat Cp = 0.56

Btu/lb°F.

Thermal conductivity k = 0.076 Btu/hr ft2°F/ft.

Viscosity µ = 0.5 cP at the average temperature of 200°F.

The tubeside fouling factor selected by the vendor is r di = 0.0015 hr ft2°F/Btu. The tubesideconfiguration is 1-in. O.D. (0.87-in. I.D.), 30 ft long, 5 rows deep, on a 2 3/8-in. triangular

tube pitch with 5/8-in. fins at 10 fins per inch.

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The vendor's design is also based on an overall heat transfer coefficient of 3.73 Btu/hr °F ft2

of extended surface, a 40°F rise in air temperature, a design air temperature of 120°F, and an

air rate of 3.2 M lb/hr. The vendor specified six 11-ft-diameter fans with a 30 hp electric

motor driver on each fan. The fan specified static head is 0.5 inH2O. The vendor's exchanger face and extended surface areas are 1200 and 160,800 ft2.

TYPICAL DESIGN DETAILS FOR AIRFIN COOLERS

Basis: 1-in. O.D. tubes

0.262 ft2/ft of bare tube R s = ax/Ao

2 3/8-in. triangular spacing on tubing 5/8-in. fins at 10 fins per inch tube with a base-to-fin surface ratio of 21.2 = R s

Depth in tube rows 3 4 5 6 7 8

Unit weight, lb/ft2 face area 66 75 80 88 97 115

Typical face velocity, ft/min 630 595 565 540 510 490

Ao/af , ft2 surface area/ft2 face area 3.80 5.04 6.32 7.60 8.84 10.0

8

Note: Table 2 is from Saudi Aramco ADP-E-001 Table 1, Pg. 103.

Table 2

Calculate the air temperature rise, air rate, surface (extended) requirements, face area, number

of fans, and horsepower for each fan to confirm the vendor's design. Since the tubeside film

coefficient for an air-cooled exchanger is calculated in the same manner as was covered in the

shell and tube section of this course (ChE 101.09), assume that the tubeside film coefficient

has been calculated and hi = 250 (Btu/hr °F ft2 of inside tube area). Confirm duty Q by using

tubeside and shellside air cooler process conditions (refer to Figure 29 at the end of the text

for nomenclature).

For the configuration given, and from Table 2 (from ADP-E-001, Pg. 103), the bundle weight

per square foot of face area is 80 lb/ft2, the recommended face air velocity (Vf ) is 565 ft/min,

the ratio of fin (ax) to base (Ao) surface is 21.2, and the ratio of bare tube surface area (Ao) to

face area (af ) is 6.32.

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Example Problem 2 (Cont'd)

Tubeside:

Q = (Wt) (Cp) (T1 - T2)

= (550,000) (0.56) (250 - 150) = 30.8 MBtu/hr

Shellside:

Q = (Ws) (Cp) (t2 - t1)

= (3,200,000) (0.242) (160 - 120) = 30.9 MBtu/hr

Cp for air from Maxwell, Pg. 88 at average temperature of 140°F.

Calculate Æte for the vendor's design, referring to Figure 6.

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X = T1 - T2

T1 - t1

= 250 - 150

250 - 120 = 100

130 = 0.77

Y = t2 - t1

T1 - T2

= 160 - 120

250 - 150

= 40

100

= 0.4

FT = 0.94

LMTD =GTTD - LTTD

ln GTTD

LTTD

= 90 - 30

ln 90

30

= 60

1.099 = 54.6

LMTD = 55

² te = (FT) (LMTD) = (0.94) (55) = 52°F

250

150

160

120

T t

T t12

1 2

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Vendor's air cooler face area is 1200 ft2

Face velocity Vf is then:

Density of air @ 70°, 14.7 psia = PAIR =

29

379

60 + 460

70+ 460

= 0.07507

Air Rate @ 70°F =

3.2 x 106 lb

hr x

hr

60 minx

1

.07507

ft3

lb= 710,415 ft

3/ min @ 70°F

Vf =

Air Rate @ 70°F and 14.7 psia

a f = Velocity over face area

Vf =

710, 415

1200 = 592 ft / min

Note: See Figure 29 on page 69 at the end of the text for nomenclature.


592 ft/min = 9.9 ft/s, which is in the middle of the acceptable range of 6 to 13 ft/s (from

text) for fan discharge velocity.

Saudi Aramco typical design detail guideline chart, ADP-E-001, Pg. 103 (see Participant

Module, Table 2), recommends a Vf of 565 ft/min for a 5-row-deep airfin cooler.

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TABLE 2

TYPICAL DESIGN DETAILS FOR AIRFIN COOLERS

Basis: 1-in. O.D. tubes0.262 ft2/ft of bare tube R s = ax/Ao

2 3/8-in. triangular spacing on tubing 5/8-in. fins at 10 fins per inch tube with a base-to-fin surface ratio of 21.2 = R s

Depth in tube rows 3 4 5 6 7 8

Unit weight, lb/ft2 face area 66 75 80 88 97 115

Typical face velocity, ft/min 630 595 565 540 510 490

Ao/af , ft2 surface area/ft2 face area 3.80 5.04 6.32 7.60 8.84 10.0

8

Note: Table 2 is from Saudi Aramco ADP-E-001 Table 1, Pg. 103.

The next step in reviewing the vendor's design is confirming the overall coefficient used by the

vendor.

Tubeside:

hi = 250 Btu/hr °F ft2 (given)

Shellside: (See Figure 29.)

Vmax

= Vf

Fraction of free flow area between tubes=

Vf

P − dR ( )P

=592

2 3 / 8 - 1( )2 3 / 8

=592

.579= 1023 ft / min

ho

=1. 9( ) dR ( ) Vmax( )0.56

P( ) Nf ( )do[ ]0.5 =

1.9( ) 1( ) Vmax( )0.56

2 3 / 8( ) 10( ) 1 + 5 / 8 + 5 / 8( )[ ]0.5

=

1. 9( ) 1023( )0.56

2.375( ) 22.50( )0.5 =

1.9( ) 48. 5( )2.375( ) 4.74( )

= 8.19 Btu / hr • F ft2

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Since • R s = ax/Ao and dr /di = Ao/ai then

ax

ai

=R s( ) dR ( )

d i

=21. 2( ) 1. 0( )

0.87= 24.37

(From Table 2, R S = 21.2.)

1

Ux

= 1

hi

ax

ai

+ r di ax

ai

+ r m + 1

ho

Ignore r m, since the value is very small.

1

Ux

=1

250

(24. 37) +(0.0015)(24.37) +

1

8.19

Ê

Ë Á

ˆ

¯˜

1

Ux

= 0.097+ 0.0366+ 0.12 = 0.25

Ux = 3.94 Btu/hr °F ft2 (extended surface)

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The calculated value for Ux (heat transfer coefficient based on square feet of extended

surface) is 3.94 Btu/hr °F ft2 as compared to the vendor-quoted value of Ux = 3.73 Btu/hr °F

ft2. Therefore, the specified extended surface 160,800 ft2 is probably adequate. However,

the equation Q = (Ux) (ax) (Æte) should be utilized to confirm the adequacy of the vendor'sspecified extended surface:

Q = (Ux)(ax)(∆ te)

31,000,000 = (3.94)(ax)(52)

a =

31, 000, 000

3.94( ) 52( )=151, 308 ft

2 (extended surface)

This calculation shows that 151,308 ft2 of extended surface is needed based on the calculatedUx [3.94 Btu/hr °F ft2 (extended)]. The vendor has specified 160,800ft2. This conclusively

shows that the vendor's specified surface is adequate.

The calculated face velocity Vf of 580 ft/min (9.7 ft/s) approximates the Saudi Aramco face

velocity recommendation for this type of air cooler, which is about 565 ft/min. Also, the 9.7

ft/s calculated velocity falls within the acceptable Vf range of 6 to 13 ft/s. Therefore, it can be

concluded that the vendor quoted face area of 1200 ft2 is acceptable.

When reviewing a vendor's design, both the tubeside and air side pressure drops must be

calculated and compared to the vendor's quoted value. Due to class time limitations, the pressure drop analysis on the tubeside will not be done as part of this example problem. The

tubeside pressure drop calculation for air coolers is done in the same manner that was

presented for the shell and tube exchangers.

Next, the fan facilities (number of fans, horsepower, and head for fans) should be reviewed.

In order to have proper air distribution over the face of the airfin exchanger, the total face area

of the fans should be 40-60% of the airfin face area.

The vendor specified 6 fans, with an 11-ft diameter for each. Calculate fan face area as a

fraction of exchanger face area:

Fan Face

Exchanger Face=

π Fan Diam2

4 x No. Fans

af

=

π 112

4 x 6

1200= 0.48

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The ratio of fan face area to exchanger face area is 0.48, which is suitable since the acceptable

range is 0.4 to 0.6.

(Tube side pressure drop; the same calculation as for shell and tube exchangers.)

Next, calculate the air side pressure drop and compare the calculated value with the fan static

head that the vendor has specified.

Fan velocity (face)Vo =Ws

(60) (fan area) (ρair ) =

3,200,000

(60) (570) (0.07507)

Vo = 1248 ft/min

PT = Pst + Pv (See Figure 9.)

Pst = (number of tube rows) (²P)

From Figure 9, Pv = 0.097.

conversion of air velocity to velocity pressure

Vo = air velocity for 70°F air and 29.92 inHg pressure, ft per min

Pv = velocity pressure, in H2O

Vo Pv Vo Pv Vo Pv

100 0.001 1,100 0.075 2,100 0.274

200 0.002 1,200 0.090 2,200 0.302

300 0.006 1,300 0.105 2,300 0.328

400 0.010 1,400 0.122 2,400 0.359

500 0.016 1,500 0.140 2,500 0.390

600 0.022 1,600 0.160 2,600 0.421

700 0.031 1,700 0.180 2,700 0.454

800 0.040 1,800 0.202 2,800 0.489

900 0.050 1,900 0.225 2,900 0.524

1,000 0.062 2,000 0.249 3,000 0.561

Figure 9

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[(ÆP) (De)] is plotted as a function of k and Vmax (See Figure 10).

k = N f ( ) do( )

dR ( )0.2

=10( )1 + 5 8 + 5 8( )

1( )0.2

=10 2.25( )

1

= 22.5

where: Nf = Number of fins/in. tube length.

do = Fin outside diameter, in.

dR = Inside fin diameter, in. (Root diameter)

De =P

d R

Nf

do

12

=

2 3 / 8

1

210( 1 + 5 / 8 + 5 / 8(

12

= (5.64)(1.88) = 10.6

where: P = Tube pitch, in.

From Figure 10:

(ÆP) (De) = 0.85 at Vmax = 1023 ft/min and k = 22.5

²P = 0.85

10.6 = 0.08 inH2O

Pst = (number of tube rows) (²P) = (5) (0.08)

= 0.4 inH2OPT = Pst + Pv = 0.4 + 0.097 = 0.5 inH2O

As discussed, the most likely head available from an air fan is 0.5 in H2O. Therefore, the

calculated value of 0.5 in H2O is acceptable.

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PRESSURE DROP CORRELATION FOR AIR

(use photostat)

Figure 10

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Calculating Plot Area

The plot area required for a horizontal air-cooled exchanger is the exchanger face area plus a

suitable buffer area between pieces of equipment on all sides of the air-cooled exchanger, to

prevent the flow of hot air from other equipment into the exchanger air supply. This buffer

area will vary, depending on the type and temperature level of the equipment upwind from the

air-cooled exchanger.

Plot area can be reduced by using an A-frame type of construction. The plot area for an A-

frame configuration is about half that required for a horizontal configuration. However, when

an A-frame configuration is used to meet plot area constraints, extra attention must be given

to the possible flow of warm air into the air supply of the A-frame exchanger. A-frame

exchangers are more susceptible than horizontal exchangers to the recirculation of their own

hot air discharge into the air supply.

Air-cooled exchangers are also very susceptible to damage from fire and maintenance

equipment because of the large area of extended surface. Therefore, the refinery space guide

should be consulted for the proper buffer area and the recommended limitation on what

equipment can be under or over an air-cooled exchanger. A sample equipment spacing guide

is shown in Figure 11. To use this guide, enter the chart from the equipment being located,

read down to the row concerning the pieces of equipment nearby, and read the minimum

distance required between them.

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ONSITE SPACING CHART

Figure 11

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OTHER TYPES OF HEAT EXCHANGERS

General

A number of secondary uses for heat-transfer equipment have not been discussed in the

preceding modules of this course. See Figure 12 for a typical listing. Some of the more

frequently utilized exchanger types will be discussed in this section of ChE 101.10. Note that

not all types of exchangers will be discussed, rather, only the more common ones.

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OTHER HEAT EXCHANGER TYPES

Type Major Characteristics Application

Double Pipe Pipe within a pipe; inner pipe may befinned or plain. For small units.

Extended Surface Externally finned tube. Services where the outside tube resistance

is appreciably greater than the inside

resistance. Also used in debottlenecking

existing units.

Brazed Plate Fin Series of plates separated by corrugated

fins.

Cryogenic services; all fluids must be

clean.

Spiral Wound Spirally wound tube coils within a shell. Cryogenic services; fluids must be clean.

Scraped Surface Pipe within a pipe, with rotating bladesscraping the inside wall of the inner pipe.

Crystallization cooling applications.

Bayonet Tube Tube element consists of an outer and

inner tube.

Useful for high temperature difference

between shell and tube fluids.

Falling Film Coolers Vertical units using a thin film of water in

tubes.

Special cooling applications.

Worm Coolers (Box

Coolers)

Pipe coils submerged in a box of water. Emergency cooling.

Barometric Condenser Direct contact of water and vapor. Where mutual solubilities of water and process fluid permit.

Cascade Coolers Cooling water flows over series of tubes. Special cooling applications for very

corrosive process fluids.

Impervious Graphite Constructed of graphite for corrosion

protection.

Used in very highly corrosive heat

exchange services.

Plate and Frame Heat

Exchanger

Series of parallel corrugated plates

separated by flexible gaskets. Process

fluids flow in nearly true counterflow

compact size.

Where high heat transfer effectiveness is

required.

Spiral Heat Exchanger Two long parallel plates wound in spiral

shape. Process fluids can be designed for

counterflow or cross flow.

Where high fouling process fluids or

slurries are present. Also used as lower

mounted condensers or Thermosiphon

Reboilers.

Figure 12

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Plate/Frame

A plate and frame heat exchanger (PFHE) consists of multiple grooved plates compressed

together by bolts. The liquid between the plates is contained by a gasket material compressed

between the plates. At each end of the exchanger is a header plate containing the inlet and

outlet parts. Plates can be added or removed as required for a service. See Figure 13 for an

exploded view of a plate and frame exchanger. The PFHE has the following advantages and

disadvantages compared to a conventional shell and tube heat exchangers.

Advantages:

• It can be disassembled for cleaning.

• The plates can be rearranged, added to or removed from the plate rack for different

service conditions.

• The fluid residence time is short (low ratio of fluid volume to surface area).

• No hot or cold spots exist to damage temperature sensitive fluids.

• Fluid leakage between streams cannot occur unless plate material fails.

• Fluid leakage due to a defective or damaged gasket is external and easily detected.

• Low fouling due to the high turbulence created by the plates.

• A very small plot area required relative to a shell and tube type heat exchanger for

the same service.

• The maintenance service area required is within the frame size of the exchanger.

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Disadvantages:

• Care must be taken by maintenance personnel to prevent damage to the gaskets

during disassembly, cleaning, and reassembly.

• A relatively low upper design temperature limitation exists (usually set by gasket

material at 300-400°F range).

• A relatively low upper design pressure limitation exists (for most units, this is in

the 140-230 psig range).

Gasket materials are not compatible with all fluids.

With permission from the Gas Processors Suppliers Association. Source: Engineering DataBook.

The plate and frame exchanger is generally considered to be a high-heat-transfer, high-

pressure-drop device. Of course, the plates can be arranged to give a low pressure drop, but

then the heat transfer coefficient also decreases. In a typical high-pressure-drop service, the

fluid flow through the exchanger is very turbulent and minimizes potential fouling. When

alloy materials are required for a heat exchange service, the plate and frame exchanger is

competitive with the shell and tube exchanger for conventional services.

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PLATE AND FRAME HEAT EXCHANGER

MoveableEndCover

Plate Pack

FixedEndCover

Carrying Bar Compression Bolt

Carrying Bar

Courtesy Alfa-Laval

Figure 13

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Hairpin

Another type of exchanger that is used for small duty services is the hairpin heat exchanger,

which is designed in a hairpin shape and fabricated in two types: double pipe and multitube.

Figure 14 shows hairpin double pipe and multitube exchangers.

The double pipe exchanger is manufactured with a bare or longitudinal finned tube. The

advantages and disadvantages of a hairpin exchanger are as follows:

Advantages:

• The use of longitudinal finned tubes will result in a compact heat exchanger for shellside

fluids having a low heat transfer coefficient.

• Countercurrent flow will result in lower surface area requirements for services having atemperature cross.

• Potential need for expansion joint is eliminated due to U-tube construction.

• Shortened delivery times can result from the use of stock components that can be

assembled into standard sections.

• Modular design allows for the addition of sections or the rearrangement of sections for

new services.

• Simple construction leads to ease of cleaning, inspection, and tube element replacement.

Disadvantages:

• Hairpin sections are specifically designed units which are normally not built to any

industry standard other than ASME Code. However, TEMA tolerances are normally

incorporated, wherever applicable.

• Multiple hairpin sections are not always economically competitive with a single shell and

tube heat exchanger.

• Proprietary closure design requires special gaskets.


Book.

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Box cooler

(Submerged Pipe Coil)

Source: Process Heat Transfer by Donald Q. Kern, page 724, copyright 1950 and 1978 by McGraw-HillBook Company, Inc. With permission from McGraw-Hill Book Company, Inc.

Figure 15

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Enhanced Heat Transfer

The heat transfer surface for the different types of exchangers can be improved (higher rate of

heat transfer per exchanger unit size) by enhancing the heat transfer surface. The

enhancement can be produced by:

• Adding fins.

• Providing nucleate boiling surfaces.

• Installing turbulence promoters.

• Providing online mechanical cleaning of the surface.

Sometimes, enhancement techniques are used in combination. The enhancements can be usedon an existing exchanger where more heat transfer is needed. A service duty increase or

temperature driving force decrease, resulting from changes to the processing conditions, can

justify surface enhancement. Examples of some high-performance fin geometrics and

enhanced boiling surfaces are shown on Figures 16 and 17, respectively. Figure 18

summarizes the use of special surface geometrics and other enhancement techniques for

various heat transfer modes.

When an enhanced heat transfer surface is evaluated for an application:

• The dominant thermal resistance in the exchanger must be identified.

• Exchanger service limitations such as flow rate, pressure drop, etc., must be quantified.

• The design objective defined (will Q, Uo, or A be held constant, and which variable will

be changed.)

Enhancement will not significantly change the overall Uo unless it alters the dominant

thermal resistance. Finally, practical concerns of cost and possibility of fouling must be

evaluated before the design is fabricated.

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HIGH-PERFORMANCE FIN GEOMETRICS

Enhanced surface for gases (A) to (F) and for condensing (G).

(A) Offset strip fins used in plate-fin heat exchanger.

(B) Louvered fins used in automotive heat exchangers.

(C) Segmented fins for circular tubes.

(D) Integral aluminum strip finned tube.

(E) Louvered tube-and-plate fin.

(F) Corrugated plates used in rotary regenerators.

(G) Integral low-fin condenser tube.

Figure 16

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ENHANCED BOILING SURFACES

(A) Rolled-over low fins.

(B) Tunnel and pore arrangement.

(C) Flattened low fins.

(D) Knurled low fins.

(E) Sintered porous metallic matrix surface.

Figure 17

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APPLICATION OF ENHANCEMENT TECHNIQUES

Commercial

Availability

Forced

Convection

Boiling

Mode

Conden-

sation

Typical Material Performance

Potential

Inside Tubes

Metal coatings Yes -- 2 -- Al, Cu, Steel High

Integral fins Yes 2 3 4 Al, Cu High

Flutes Yes 4 4 4 Al, Cu Moderate

Integral roughness Yes 2 3 4 Cu, Steel High

Wire coil inserts Yes 3 4 4 Any Moderate

Displaced promoters Yes 2 4 4 Any Mod. (lam)

Twisted tape inserts Yes 2 3 4 Any Moderate

Outside Circular Tubes

Coatings

Metal Yes -- 2 4 Al, Cu, Steel High (boil) Nonmetal No -- 4 4 "Teflon" Moderate

Roughness (integral) Yes 3 2 4 Al, Cu High (boil)

Roughness (attached) Yes 3 4 -- Any Mod. (for conv)

Axial fins Yes 1 4 4 Al, Steel High (for conv)

Transverse fins

Gases Yes 1 -- -- Al, Cu, Steel High

Liquids/two-phase Yes 1 1 1 Any High

Flutes

Integral Yes -- -- 2 Al, Cu High

Nonintegral Yes -- -- 4 Any High

Plate-Fin Heat Exchanger

Metal coatings Yes -- 3 -- Al High

Surface roughness Yes 4 3 4 Al High (boil)

Configured or

interrupted fins

2 2 Al, Steel High

Flutes No -- -- 4 Al Moderate

Plate Type Heat Exchanger

Metal coatings No -- 4 -- Steel Low

Surface roughness No 4 4 4 Steel Low

Configured channel Yes 1 3 3 Steel High (for conv)

where: 1 = Common use.

2 = Limited use.3 = Some special cases.

4 = Essentially no use.

Figure 18

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CALCULATING FURNACE EFFICIENCY

General

All furnaces and heaters are classified in one of two categories; direct fired heaters or fire tubeheaters. Because most furnaces and heaters in a refinery are direct fired, the following

discussion will be limited to direct fired equipment. However, a brief summary of fire tube

heater types, their characteristics, and how they compare with direct fired heaters is given

below as general background.

Direct Fired Fire Tube

Applications

Hot oil heater. Indirect fired water bath heaters (line heaters).

Regeneration gas heaters. Propane and heavier hydrocarbon vaporizers.

Amine and stabilizer reboilers. Hot oil and salt bath heaters.

Glycol and amine reboilers.

Low pressure steam generators.

Characteristics

More ancillary equipment and controls. Heat duty usually less than 10 MBtu/hr.

Higher thermal efficiency. Easily skid mounted.

Requires less plot space. Forced or natural draft combustion.

Forced or natural draft combustion. Less likely to have hot spots or tube rupture.


Book.

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Direct Fired

There are two basic types of direct fired furnaces, cylindrical and cabin. Within each type there

are many different configurations. The furnaces can have different coil arrangements:

horizontal, vertical, helical, or serpentine. Also, the furnace can be all radiant (no convection

section) or have a convection section. Several configurations for the vertical cylindrical and

cabin type furnaces are shown in Figures 19 and 20.

The all radiant cylindrical furnace is the simplest and least expensive. Typically, an all-radiant

furnace operates with about a 60% efficiency and a stack temperature of about 1200°F. Adding

a convection section to an all-radiant vertical cylindrical furnace increases the overall furnace

efficiency to about 80%. Of course, the convection section significantly increases the furnace

cost.

Some of the advantages for the two types of direct-fired furnaces are as follows.

Cylindrical Furnace Advantages:

• Require the smallest plot area for a given duty.

• The cost is usually 10 to 15% lower in the larger sizes.

• Can accommodate more parallel passes in the process coil.

• For large duties, a cylindrical heater has a taller firebox and more natural draft at the

burner.

• The flue gas velocity is usually higher in the convection section, hence, the flue gas film

coefficient is higher.

• Few expensive tube supports or guides are required in the convection section.

• The noise plenums or preheated combustion air plenums are smaller.

• Fewer soot blowers are required in the convection section. (Soot blowers are not needed

for gaseous fuel.)

• If coil drainage is a problem, a helical coil may be used when there is only one pass.

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Cabin Furnace Advantages:

• The process coil can always be drained.

• Two-phase flow problems are less severe (slug flow can generally be avoided).

• Cabins can accommodate side-firing or end-firing burners instead of only vertically

upward firing. This permits the floor of the heater to be closer to the ground. (Some

burner manufacturers prefer to fire liquid fuels horizontally.)


Book.

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EXAMPLES OF VERTICAL CYLINDRICAL DIRECT FIRED FURNACES


Figure 19

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EXAMPLES OF CABIN DIRECT FIRED FURNACES


Figure 20

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The major components of a furnace are the radiant section (firebox), convection section,

stack, burner fuel system, and process fluid coil. The radiant section provides the high level

heat to the process coil, with the process fluid flow leaving the furnace via the radiant section.

The burner flame is contained in the radiant section. The combustion gases leaving the

radiant section typically are in the 1500-1900°F temperature range. Heat is transferred from

the flame to the process coil mainly by radiation from the flame.

The hot gases of combustion leave the radiant section and flow into the convection section,

which transfers the low-level heat to the cold process fluid as it enters the furnace. The

combustion gases are cooled in the convection section from the 1500-1900°F range to less

than 750°F. Heat is transferred from the gases of combustion (flue gas) to the process fluid

coil via convection (hot gas moving over pipes).

The stack sits on top of the convection section and generates sufficient draft to overcome the

friction losses of the hot flue gas flowing over the convection section tubes. If pollutionconsiderations set the stack height higher than is needed for draft, a damper in the stack can

absorb the incremental available draft. Refer to Figure 21 for an illustration of furnace draft.

The burner/fuel system includes the burner, which mixes air with fuel and burns the fuel in

the radiant section of the furnace. The burner flame typically is about 60% of the height of

the radiant section. Fuel and air are fed to the burner by separate pipe/duct systems.

The process fluid coil carries the process fluid being heated in the furnace from the process

inlet in the convection section (flue gas outlet) to the process outlet in the floor of the radiant

section. The coil changes in configuration (horizontal, vertical, low tube finned tube) and

type of materials throughout the furnace. The coil is exposed to relatively mild conditions atthe process inlet in the convection section and to severe conditions in the radiant section.

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FURNACE NATURAL DRAFT PROFILES


Figure 21

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The fuel to be fired required to meet the QA duty is further adjusted for heat loss from the

furnace firebox (radiant section). This loss typically is about 2%. Therefore, the gross

quantity of fuel fired to meet QA duty is FG = (F N) (1.02). QF can be determined from the

equation QF = FG (lb/hr) x LHV (Btu/lb). Furnace efficiency is:

Percent efficiency =QA (100)

QF

The following backup calculation can be done to check the furnace efficiency calculation.

Fuel to a furnace is measured by a flowmeter. The actual rate of fuel should be determined

from the fuel meter and a backup of QF value calculated from the fuel meter reading. If there

is a significant disagreement between QF calculated from the efficiency equation and QF

calculated from the fuel meter, this difference should be reconciled before the calculated

furnace efficiency is accepted as a credible value.

The percent excess air at the burner is calculated from the furnace flue gas analysis. The

quantity of excess air at the burner affects the amount of energy in the fuel that is available for

absorption by the process fluid coil.

Example Problem 3

Part 1 - Calculate Percent Excess Air

The measured flue gas composition is (all values vol% for dry flue gas) CO2 = 9.5,

CO = 1.8, O2

= 2.0 and N2

= 86.7. The total oxygen to the furnace is determined from the

quantity of N2 in the flue gas and the fact that air is 79 vol% N2 and 21 vol% O2.

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Example Problem 3, Part 1 - Percent Excess Air

The measured flue gas composition is (all values vol% for dry flue gas) CO2 = 9.5 CO = 1.8 O2

= 2.0 and N2 = 86.7. The total oxygen to the furnace is determined from the quantity of N2 inthe flue gas and the fact that air is 79 vol% N2 and 21 vol% O2.

O2 supplied to furnace

100 moles of flue gas =

moles of N2

100 moles of flue gas moles of O2

100 moles of air

moles of N2

100 moles of air

(86.7) (21)

(79) = 23.05 moles of O2

100 moles of flue gas

Excess oxygen at the burner is expressed on the basis of complete combustion; namely, all

hydrogen and carbon in the fuel are burned to water and carbon dioxide. Therefore, the

amount of free oxygen in the flue gas analysis must be reduced by the amount needed to

complete the combustion of the flue gas CO to CO2. The formula for this adjustment is:

(flue gas CO content) + 0.5 moleO2

mole CO = CO2

1.8 moles CO

100 moles flue gas + 0.9 mole O2

100 moles flue gas = 1.8 moles CO2

100 moles of flue gas

Adjusted excessO2 = 2.0 molesO2100 moles of flue gas

- 0.9 moleO2100 moles of flue gas

= 1.1 molesO2100 moles of flue gas

From the flue gas analysis and these calculations, the excess oxygen at the burner can be

determined as follows:

Percent excess O2 =

adjusted moles excess O2

100 moles flue gas (100)

(moles O2 burnt in furnace)

100 moles flue gas

=


100 moles flue gas (100)

(moles O2 supplied to furnace)

100 moles flue gas -


100 moles flue gas

=(1.1) (100)

23.05 - 1.1 = 110

21.95 = 5%

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Since the percent excess O2 is the same as the percent excess air, the sample calculation

shows that the furnace under study is operating with 5% excess air. The nomogram in Figure22 cqn also be used to determine percent excess air after the adjusted percent oxygen in the

flue gas has been calculated.

FLUE GAS OXYGEN VERSUS EXCESS AIR

(drawing)

Figure 22

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Example Problem 3, Part 2 - Furnace Efficiency

Calculate the thermal efficiency for a furnace where the operating duty QA has been

confirmed to be 150 MBtu/hr [specific heat equation was used to confirm duty, Q = (W)(Cp)(Dt)]. Also, the flue gas temperature is 700°F, the flue gas composition is the same as in

Part 1 (use 5% for excess air at burner), heat losses from firebox are 3%, and fuel gas fired in

the furnace has a LHV of 19,700 Btu/lb.

Net fuel = F N = AHA

= 150,000,00016,600

= 9040 lb/hr fuel (See Maxwell, Pg. 184 forHA)

Gross fuel = FG = (F N) (firebox heat loss) = (9040) (1.03) = 9310 lb/hr fuel

Heat fired = QF = (FG) (LHV fuel) = (9310) (19,700) = 183,350,000 Btu/hr

Percent efficiency=Q A( )100( )

QF

=150,000,000

183,350,000= 82%

For a quick approximation of furnace efficiency, the following shortcut formula can be used

in conjunction with Figure 22 (percent excess air versus percent O2 in flue gas curve).

Percent efficiency

(LHV) = [(100 - (0.0237 + (0.000189) (EA))) (TST - TA)] 100

100 + QL

where: EA = Percent excess air.

TST = Stack temperature, °F.

TA = Ambient air temperature, °F.

QL = Casing heat loss.

For Example 3, Part 2 conditions, the furnace efficiency calculated from the shortcut formula

is as follows:

Percent efficiency = [(100 - (0.0237 + (0.000189) (5))) (700 - 80)] 100

100 + 3

Percent efficiency = [100 - (0.0246) (620)] (0.971) = 82.3

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Routine Furnace Startup and Operations

Startup

The complexity of fired heaters is increasing. Today, furnace complexity often dictates that afurnace startup advisor be present for major startups. The advisor, together with mechanical,

instrument, and burner specialists, review in detail the heater piping and instrumentation.

Upon completion of this review and corrective actions required, the heater is ready to be lit

for lining dryout. The following activities are expected from the startup personnel during the

lining dryout and initial furnace operation.

• Review the Operating Manual and revise the fired heater section as necessary (prestartup,

oil in, normal operations, shutdown procedures, troubleshooting, and auxiliary equipment

instructions).

• Ensure that hydrostatic test water has been removed from the coil to the maximum extent practical.

• Ensure that all fuel lines have been steam blown (not through the burner guns).

• Check the performance of all the burners during refractory dryout.

• Monitor thermal movements of tubes, tube support systems, and refractory during dryout.

Watch for debris on the heater floor.

• Investigate any performance data for the fired heater and attendant equipment, that

appears to differ from design specification values. Listed below are some of the moreimportant general observations to be made and problems to look for during an initial

startup.

_ Coil and external piping movements.

_ Lining condition as heater reaches operating temperature.

_ Pass flows, pass crossover, and outlet temperature.

_ Tube hot spots and overheated passes.

_ Burner and pilot combustion performance. Watch for problems such as fuel

dripback, gun orifice plugging, gun tip coking, wet atomizing steam, uneven

burner firing rates, leaning flames, flame impingement, burner noise, etc.

_ Draft conditions, particularly at bridgewall. _ Combustion air pressure.

_ Expansion joint movement.

_ Damper positions.

_ Stack vibration.

_ Fan-induced vibration and noise.

_ Unsafe operating practices.

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• Discuss special problems related to the specific fired heater in the operating manual.

When all prestartup activities have been completed (equipment checkout completed, linings

dried, etc.), at the appropriate time in the unit oil-in operation, the furnace will be lit and put

online. The following furnace startup steps are listed for background information and should

not be considered complete. In each startup procedure, certain aspects of the procedure are

unique to a particular service.

• Check to see if all fuel and pilot systems are active up to unit battery limits.

• Check to see if all drains and vents of the on-fuel/pilot systems are closed.

• Check to see if all instrumentation is working and that automatic shutdown devices are

deactivated.

• Commission any fuel oil steam tracing and open blinds at battery limits on fuel/pilot

supplies.

• Steam out fuel oil system to bring piping up to temperature.

• Open furnace stack damper fully.

• Start snuffing steam to furnace firebox and shut off snuffing steam when a good flow of

steam can be observed from the stack.

• Fully open air dampers on each burner.

• Open valves on pilot gas system to purge inert gases.

• Ignite fuel gas pilots, one burner at a time.

• Open valve to bring fuel oil and tracing steam into the burner supply systems.

• Start atomizing steam to the first burner.

• Slowly open the fuel oil valve to the first burner and observe ignition. Adjust oil and air

rates to give a stable, nonluminous flame. Set firing at a minimum stable rate consistentwith a good flame pattern.

• Repeat for each burner.

• When all burners are lit, check for proper operation of pilot and burner flames.

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• Activate furnace instrumentation and raise furnace coil outlet temperature at the rate of

about 50°F/hr.

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Optimum Excess Air Levels

As part of the discussion on furnace efficiency, some of the furnace operating variables have

been discussed, namely:

• Checking the furnace duty QA by the enthalphy heat equation QA = W(Æh), and

• Utilizing the stack flue gas temperature and oxygen content as part of burner operating

conditions review.

For these variables operating the burner with the correct amount of excess air (determined

from O2 level in flue gas) has the most significant effect on the entire operation of the

furnace. Therefore, this discussion will further explore proper excess air levels for furnace

burners.

Figures 23 and 24 show the effect of different levels of excess air on the furnace efficiency

and level of combustibles in the stack flue gas. An excess air target should be established for

each furnace, and the operating level versus the target level should be monitored by the plant

engineer or by automatic instrumentation.

The target excess air level is established by plugging air leaks in the furnace walls and then

reducing the air rate to the burner in increments while monitoring the carbon monoxide and

smoke level in the stack flue gas. When the carbon monoxide level reaches the 100-200 ppm

range, the minimum acceptable excess air level has been reached. The actual monitored

target level for excess air will be at or close to this minimum level, as determined by theactual furnace service and associated instrumentation under study. Significant fuel savings

can be made by monitoring burner excess air: a 40°F decrease in flue gas temperature usually

produces about a 1% increase in furnace efficiency. The cost effect of unplugged air leaks is

shown in Figure 25. Lowering the stack temperature to improve efficiency is usually limited

by return on investment and the acid dew point in the flue gas (discussed in ChE 101.02).

Operating Guidelines

Low Draft High Draft

Low Excess Air (O2) Open Damper Open Burner Air

High Excess Air (O2) Close Burner Air Close Damper

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OPTIMUM EXCESS AIR FOR A FIRED HEATER

Figure 23

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TYPICAL COMBUSTIBLES EMISSION FROM FIRED HEATERS

Figure 24

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COST OF FURNACE AIR LEAKS

(use photostat)

Figure 25

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EXAMPLE OF DIRECT FIRED REBOILER

CONTROLS/SAFETY DEVICES


Figure 26

The alarms and a description of the shutdown systems shown in Figure 26 are listed on the

following page.

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HEATER ALARM/SHUTDOWN SYSTEM DESCRIPTION

Caution: The alarms and shutdowns shown do not necessarily meet any minimum safety

requirement, but are representative of the types used for control systems.

Basic Criterion: The failure of any one device will not allow the heater to be damaged.

Schematic

Label

Alarm/Shutdown

Description

Regeneration

Gas Heater

Hot Oil Heater and

Direct Fired Reboiler

TSH-1 High stack temperature. See Note 1. See Note 1.

TSH-2 High outlet temperature. See Note 1.

FSL Low mass flow through tubes. See Notes 2 and 4. See Notes 3 and 4.

BSL Flame failure detection. See Note 5. See Notes 5 and 6.

PSL Low fuel pressure. See Note 6.

PSH-1 High fuel pressure. See Note 7. See Note 7.

PSH-2 High cabin pressure. See Note 8. Not applicable, if

natural draft.

Notes:

1. A direct immersion jacketed thermocouple is preferred because the response is ten timesfaster than a grounded thermocouple in a well. A filled bulb system is a poor third choice.

The high stack gas temperature shutdown should be set approximately 200°F above

normal operation.

2. An orifice plate signal should be backed up by a low pressure shutdown to ensure

adequate process stream flow under falling pressure conditions.

3. The measurement should be on the heater inlet to avoid errors from two-phase flow.

4. Differential pressure switches mounted directly across an orifice plate are not satisfactory

due to switch does not turn on at the same pressure as it turns off. An analog differential pressure transmitter with a pressure switch on the output is recommended. The analog

signal should be brought to the shutdown panel so that the flow level can be readily

compared with the shutdown point. With permission from the Gas Processors Suppliers

Association. Source: Engineering Data Book.

5. The flame scanner should be aimed at the pilot so that a flameout signal will be generated

if the pilot is not large enough to ignite the main burner.

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6. If the heater design precludes flame scanners, a low fuel gas pressure shutdown should be

installed to prevent unintentional flameout. This shutdown should detect gas pressure at

the burner.

7. Either burner pressure or fuel control valve diaphragm pressure may be used. This

shutdown should be used whenever large load changes are expected. It prevents the

heater from overfiring when the temperature controller drives the fuel wide open to

increase heat output with insufficient air.

8. This shutdown should block in all lines to the heater because the probable cause of its

activation is tube rupture. Gas is probably burning vigorously outside the heater.

9. With permission from the Gas Processors Suppliers Association. Source: Engineering

Data Book.

The parts of the fired heater instrumentation related to safety should be given special attentionand regularly inspected as well as tested for functionality. The following items, either

observed by the plant engineer or indicated by instrumentation, usually indicate a problem

with furnace operations.

• The burner flame is not symmetrical, pulsates or breathes, is unusually long or lazy, lifts

off the burner, etc.

• The burner is not aligned and/or the flame is too close to the tubes.

• There is a lack of negative pressure (draft) at the top of the firebox.

• The stack gas is smokey.

• The gas in the firebox appears hazy.

• There are unequal temperatures, differing by more than 10°F, on the process pass outlets.

• The stack temperature increases steadily with no change in the process heat duty.

• The fuel gas control valve is wide open.

• The fuel gas composition or pressure varies widely.

• The tubes in the heater are not straight.

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Book.

Burner selection is very important because an improper burner will reduce furnace efficiency

and service factor. Plant personnel need to have general knowledge of burners so that:

• The proper maintenance will be performed.

• Burner operating problems can be properly diagnosed and corrected.

• The burner operation can be optimized.

Four types of burners are commonly used in direct fired heaters.

• Inspirating Pre-Mix Burners - The passage of fuel gas through a venturi pulls in the

combustion air. These burners have short dense flames that are not affected by wind

gusts.

• Raw Gas Burners - Some of the air required for combustion is pulled in by a venturi. The

rest of the air is admitted through a secondary air register. These burners have larger

turndown ratios, require lower gas pressures, and are quieter.

• Low NOx Burners - The addition of a tertiary air register reduces the amount of nitrogen

oxides in the flue gas. This type also can be operated with less excess air than inspirating

pre-mix or raw gas burners.

• Combination Gas and Oil Burner - An oil burner is added to the gas spider so that fuel oilcan also be used. One-tenth pound of steam per pound of fuel is usually required to

atomize the oil.


Book.

Figure 27 shows a cross-sectional view of a combination gas/oil burner. This module

(ChE 101.10) contains limited information on the selection and operation/maintenance of

burners. Future courses will explore these subjects in greater depth.

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NATURAL DRAFT OIL/GAS BURNER

Figure 27

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Monitoring Tube Metal Temperature

One critical variable to monitor in many furnaces after startup is the tube metal temperatures

(temperature of the process coil on the firebox side) in the radiant section. Tube failures

account for more than half of the furnace fires and explosions. Excessive tube metal

temperatures accelerate tube creep (sagging tubes), hydrogen attack, and external (vanadium

attack, oxidation) and internal corrosion of the tubewall. Monitoring tube metal temperatures

also helps define the end of the current furnace run, the point at which the furnace is due for a

shutdown and decoking. Tube metal temperatures increase with coke laydown in the tube,

assuming all other variables are held constant. Tube metal temperatures are monitored by

thermocouples attached to the tube, or by a pyrometer that measures the radiation emitted by

the furnace tubewall. An expert on tubewall temperature instruments should be consulted

about the installation of thermocouples or the purchase of a pyrometer. Both measuring

devices are sophisticated pieces of equipment that vary in type depending on the service.

Usually, the highest tube metal temperature in a direct fired heater occurs in the radiant

section, where the process fluid temperature on the inside of the tube is the highest. The

maximum allowable operating tube metal temperature for any one tube material is a function

of the tubewall stress level and the severity of the tubewall corrosive atmosphere. Figure 28

gives allowable stress levels for furnace tubes as a function of wall temperature for several

material types.

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ALLOWABLE ELASTIC AND CREEP RUPTURE STRESS

FOR TYPICAL HEATER TUBE MATERIALS

(Elastic and Creep Rupture Stress, psi)

Medium

Carbon Steel C-1/2 Mo 1-1/4 Cr-1/2 Mo 2-1/4 Cr-1 Mo 5 Cr-1/2 Mo

Temp.

°F(1)

Elastic

Stress

Creep

Stress

Elasti

c

Stress

Creep

Stress

Elastic

Stress

Creep

Stress

Elasti

c

Stress

Creep

Stress

Elasti

c

Stress

Creep

Stress

700 15,800 20,800 15,700 15,250 18,000 16,800

750 15,500 16,900 15,400 15,000 18,000 16,500

800 15,000 13,250 15,000 14,600 17,900 15,900

850 14,250 10,200 14,500 14,250 17,500 15,200

900 13,500 7,500 14,000 17,000 13,800 17,500 17,100 16,700 14,400 13,250

950 12,600 5,400 13,400 10,250 13,300 10,900 16,500 12,100 13,500 9,6001,000 11,500 3,700 12,700 5,900 12,800 6,700 15,750 8,700 12,400 7,000

1,050 11,900 3,400 12,100 4,150 14,750 6,400 11,300 5,100

1,100 10,900 2,000 11,400 2,600 13,600 4,600 10,250 3,700

1,150 12,300 3,150 9,200 2,700

1,200 10,700 1,750 8,200 1,950

1,250

Notes:

(1) For intermediate temperatures, stresses can be obtained by graphical interpolation.

Source: API Recommended Practice 530, Calculation of Heater Tube Thickness in PetroleumRefineries, Third Edition, September 1988. Reprinted courtesy of the American Petroleum

Institute.

Figure 28

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R Overall resistance to heat flow, hr ft2 °F/Btu.

r di Fouling resistance inside tube, hr ft2 °F/Btu.

r do Fouling resistance outside tube, hr ft2 °F/Btu.

r m Resistance to heat flow of tube wall, hr ft2 °F/Btu.

R S Outside surface area finned tuOutside surface area bare tub

T1 Inlet temperature of fluid to be cooled, °F.

T2 Outlet temperature of fluid to be cooled, °F.

t1 Inlet air temperature, °F.

t2 Outlet air temperature, °F.

Æte Effective temperature difference, °F.

Ætm Logarithmic mean temperature difference (LMTD), °F.

Uo Overall heat transfer coefficient (related to bare tube O.D.), Btu/hr

ft2 °F.

Ux Overall heat transfer coefficient (related to finned tube), Btu/hr ft2

°F.

Vf Face velocity of cooling air, ft/min at 70°F, 29.92 inHg.

Vo Air velocity at outlet of fan (70°F, 29.92 inHg), ft/min.

Vmax Air velocity through the minimum free flow area of tube banks.

Ft/min of air at 70°F, 29.92 inHg.Ws Air rate, lb/hr.

Wt Process fluid rate, lb/hr.

X (T1 - T2)/(T1 - t1) (see Figure 5).

Y (t2 - t1)/(T1 - T2) (see Figure 5).

Figure 29

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KEY FORMULAS

The following is a brief summary of the more important and useful formulas in this module.

Note that many of them can be easily programmed for personal computer use.

LMTD correction factorFT: (Figure 5)

X =T1 - T2

T1 - t1Y =

t2 - t1T1 - T2

where: T1 = Inlet temperature of fluid to be cooled, °F.

T2 = Outlet temperature of fluid to be cooled, °F.

t1 = Inlet air temperature, °F.

t2 = Outlet air temperature, °F.

LMTD:

LMTD = GTTD - LTT

ln GTTDLTTD

where: GTTD = Greater temperature differen LTTD = Lesser temperature differenc

Air fin face velocity,Vf

Vf = r a r

60 29379

0.98 = ft/mi

Air velocity through min free flow area,Vmax:

Vmax =Vf

(P - dR ) af

Paf

= ft/mi

where: Vf = Face velocity, ft/min.

P = Center-center of tubes, in.

dR = Inside diameter of fins, in.

af = Face area, ft2.

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Outside film heat transfer coefficient,ho:

ho =1.9 dR (Vmax)0.56

P Nf (do)0.5

= Btu/hr ft2 °

where: Nf = Number of fins/in. of tube.

do = Outside diameter of fins, in.


P = Center-to-center distance (pitch) of tubes, in.

Vmax = Air velocity through the minimum free flow area of tube banks. Ft/min

of air at 70°F, 29.92 inHg.

Ratio finned surface to bare surface,ax/ai:

axai

=R S dR

di

where: ax = Tube finned surface, ft2.

ai = Tube inside area, ft2.

R S = Outside surface area finned tubeOutside surface area bare tube

.


di = Inside diameter of tube, in.

Overall heat transfer coefficient,Ux:

1Ux

= 1hi

ax

ai + r di

ax

ai + r m + 1

ho

where: ax = Tube finned surface, ft2.

ai = Tube inside area, ft2.

r di = Inside tube fouling resistance.

r m = Tubewall resistance.

hi

= Inside tube film coefficient.

ho = Outside film heat transfer coefficient for finned tube, Btu/hr ft2 °F.

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WORK AID 1 - PROCEDURES FOR CALCULATING PERCENT OF DESIGN DUTY

To calculate the percent of design duty an airfin exchanger can perform when the air

temperature rises above design, use the following procedure:

A.

Step 1: Calculate X and Y at design conditions:

X =T1 - T2

T1 - t1Y =

t2 - t1T1 - T2

Step 2: Determine FT, using Figure 6.

Step 3: Calculate LMTD:

GTTD - LTTD

lnGTTDLTTD

or, use chart from TEMA Manual, Pg. 11

Step 4: Calculate Æte:

Æte = FT (LMTD)

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B. At the heat wave conditions, use a trial-and-error calculation where the air temperature use

is affected by the decrease in the exchanger's capability to transfer heat. For the first trial

calculation, estimate that the duty transferred by the exchanger is decreased by 15%.

Also, since the exchanger surface is constant, the change in duty versus air temperature

must be considered. Use the formula:

Q1

U1 ∆te1

= A1 = A2 =Q2

U2 ∆te2

AssumeU1 = U2

Therefore,Q1

∆te1

=Q2

∆te2

Step 1: Calculate the process fluid temperature drop at the assumed 85% duty.

Step 2: Calculate the air temperature rise at the assumed 85% duty.

Step 3: Calculate the process and air outlet temperatures.

Step 4: Determine X and Y values.

Step 5: Determine FT, using Figure 6.

Step 6: Determine LMTD.

Step 7: Calculate Æte.

Step 8: Check the assumption of 15% heat duty reduction, using:

Q2

Q1 =

∆T2

∆T1 =should equal the assumed duty

(in this case, 0.85)

If not, assume a new duty and recalculate.

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WORK AID 2 - PROCEDURES FOR CALCULATING EXTENDED SURFACE AND

FACE AREA REQUIREMENTS

To calculate the extended surface and face area requirement of an airfin exchanger:

Step 1: Calculate the estimated air temperature use, using the formula:

Estimated use =Ux + 1

10 T1 + T2

2 - t1

Step 2: Calculate X and Y:

X = T1 - T2

T1 - t1Y = t2 - t1

T1 - T2

Step 3: Using Figure 6, determine FT.

Step 4: Calculate LMTD.

Step 5: Calculate Æte:

Æte = (LMTD) (FT)

Step 6: Calculate ax, using the formula:

ax =Q

Ux ∆te

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Step 5: Calculate net fuel fired, F N:

F N =QA

F N

Step 6: Calculate gross fuel fired, FG. Assume furnace heat losses are 2 1/2%.

FG = 1.025 (F N)

Step 7: Calculate heat fired, QF, Btu/hr.

QF = (FG) (LHV fuel)

Step 8: Calculate furnace efficiency:

% efficiency =QA (100)

QF

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GLOSSARY

acid dew point The temperature at which an acidic component in the flue gas

condenses.

A-frame exchanger Two sections of air coolers are hooked together with each

sloping 45° in opposite directions, forming an A shape. This

is done to reduce plot space.

air plenum

chamber

An empty chamber under or over an air cooler that allows air

to spread or be collected from the air cooler face area.

battery limits The real estate boundary line denoting the start of a

processing area.

bubble point curve Denotes the family of pressure/temperature points at which

the fluid starts to vaporize.

burner registers The openings, equipped with a regulating device, in the

burner to feed air to the burner.

contamination The quality of a substance is made unacceptable by a

contaminant. This reduction in quality is called

contamination.

corbeled wall An irregular wall in the convection section of a furnace. Theirregularities in the wall match the staggered tubes in the

convection section to prevent flue gas from bypassing around

the tubes.

cross flow

exchanger

The tubes carrying the process fluid in an air-cooled

exchanger usually run horizontally. The air flow is vertical;

therefore, it flows in a cross flow manner across the tubes.

dew point curve Denotes the family of pressure/temperature points at which

the fluid starts to condense.

embedded type fin The fin is wrapped around the tube, sometimes held in place

by grooves cut in the tube.

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REFERENCES

1. Chemical Engineer's Handbook, 6th Edition, R. H. Perry and D. Green (Physical

properties, general information, calc equations).

2. Data Book on Hydrocarbons, J. B. Maxwell (Hydrocarbon physical properties).

3. Engineering Data Book, Gas Processors Suppliers Association, 10th Edition, 1987.

4. Process Heat Transfer, Kern, 1950.

5. Standards of the Tubular Exchanger Manufacturers Association (TEMA), 7th Edition,

1988.

6. AES-E-001, Basic Design Criteria for Unfired Heat Transfer Equipment.

7. ADP-E-001, Exchangers.

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APPENDICES

AIR COOLED HEAT EXCHANGER SPECIFICATION SHEET

(photostat)

Figure 30

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AIR COOLED HEAT EXCHANGER SPECIFICATION SHEET (CONT'D)

(photostat)

Figure 30 (cont'd)

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(photostat)

Figure 30 (cont'd)

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(photostat)

che 10110

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