FWC air cooled exchanger.pdf

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PROCESS PLANTS DIVISION

PROCESS STD 302 PAGE 1 REV 10 DATE JULY 2002

HEAT TRANSFER AIR COOLED EXCHANGERS

INDEX Page 1.0

GENERAL

4

2.0

ECONOMICS 2.1 Air Versus Water Cooling 2.1.1 Advantages of Cooling with Air 2.1.2 Disadvantages of Cooling with Air 2.1.3 Effect of Approach Temperature 2.1.4 Cost Comparison 2.2 Air-Fin Optimization 2.2.1 Selection of Tube and Fin Dimensions 2.2.2 Number of Tube Rows 2.2.3 Design Air Velocity

5 5 5 5 6 7 8 8 9 9

3.0

PROCESS SPECIFICATION 3.1 Tubeside 3.2 Air Side

10 10 13

4.0

MECHANICAL DESIGN FEATURES 4.1 Induced or Forced Draught 4.1.1 Induced Draught 4.1.2 Forced Draught 4.2 Air Humidification 4.3 Tube Bundle Orientation 4.4 Header Type 4.5 Tubes and Fins 4.5.1 Tube and Fin Dimensions 4.5.2 Fin Types 4.6 Fans and Drivers 4.6.1 Fans 4.6.2 Drivers and Speed Reducers 4.7 Materials of Construction 4.7.1 Header, Cover, Plugs and Gaskets 4.7.2 Tubes and Fins 4.7.3 Fans 4.7.4 Corrosion Allowance 4.7.5 Design Pressure and Temperature 4.8 Distribution Piping

17 17 17 18 18 18 19 20 20 20 22 22 22 23 23 23 23 24 24 24

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INDEX (Cont’d) Page 5.0

OPERATION AND CONTROLS

27

5.1 5.2 5.3 5.4 5.5 5.6

27 27 27 28 28 29

Multiple Fans Process Bypasses Adjustable Louvers Variable Pitch Fans Air Recirculation Variable Speed Motors

6.0

PLOT PLAN

30

7.0

NOISE

32

8.0

ESTIMATION OF EQUIPMENT SIZE AND UTILITY CONSUMPTION

33

8.1 8.2

33 36

Calculation Procedure Sample Problem

Appendix A. Calculation of Design Air Temperature by a statistical method

42

INDEX OF TABLES Table No. 1 2 3 4

Title Allowable Pressure Drop Typical Overall Heat Transfer Coefficients Bundle Face Area Ratios Typical Face Velocities

Page 12 39 41 41


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INDEX OF FIGURES Figure. No. 1 2 3 4 5 6 7 8

Title

Page

Air Cooled Exch. Process Spec. Sheet (FWC Lvon) Air Cooled Exch. Process Spec. Sheet (FWC London) Overall Heat Transfer Coefficient vs. Liquid Viscosity Temperature Corr. Factor for Cross Flow Approximate Number of Tube Rows Reciprocal of Air Density (d), vs. Air Temp. and Altitude Air Side Static Press. Drop for 3 Rows Finned Tubes Correlation of Air Flowrate and Fan Diam. Vs. Vel. Press.

15 16 44 45 46 47 48 49


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1.0



GENERAL Air cooled exchangers have been in use in the petroleum industry since the early 1930's though until about 1950 their use was confined to locations where cooling water supplies were virtually non-existent. Their widespread adoption since the latter date has primarily been because of economic advantages. However, the use of water coolers has always entailed the possibility of contamination of rivers and ground waters. Increasing awareness of authorities with regard to environmental pollution will result in more air coolers being selected due to their relative freedom from environmental problems. There are few restrictions on the use of air coolers and more recently services have included cooling/condensing of turbine exhausts and condensing of vacuum tower overheads. Generally, the only services for which air coolers have not been considered, because of the relative difficulty of cleaning tubes in situ, is either where frequent cleaning of the process side is necessary, such as in the food industry, or where the process fluid may gel on standing.


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2.0



ECONOMICS 2.1

Air Versus Water Cooling 2.1.1

Advantages of Cooling with Air The advantages of direct cooling with air as compared to cooling with water in shell and tube exchangers can be summarized as follows:

2.1.2

1.

Elimination of water circulation system and possible water cooling tower installation.

2.

The problem of temperature and chemical pollution of water resources is eliminated.

3.

Exchanger maintenance is minimized as there is generally little fouling on the coolant side and this causes few problems.

4.

In the event of power failure some advantage can be taken of cooling due to natural draught and radiation.

5.

Elimination of water treatment costs and provisions to protect against freezing of coolant.

6.

Location of plant and coolers is facilitated because of independence from water source.

7.

Cooling water often requires the use of exotic materials, e.g. Admiralty Brass, Cu-Ni, Monel, etc.

8.

Air is free.

Disadvantages of Cooling with Air 1.

Air fins are more expensive.

2.

Air fins generally have a smaller MTD available.

3.

Leaking tubes may cause explosion hazards or fire.

4.

More electrical equipment is required.

5.

Air being more liable to temperature fluctuation than water, process control instruments are often required. FOSTER WHEELER ENERGY LIMITED 2002

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2.1.3



Effect of Approach Temperature The selection of air-cooling may ultimately be based on economic factors (see Section 2.1.4) taking into account the advantages and disadvantages listed in Sections 2.1.1 and 2.1.2. The approach temperature defined as the difference between the process fluid outlet temperature and the design air temperature is the major design variable, which affects the economics. The cost of an air fin and its ancillary equipment, assuming a fixed duty, will decrease with increase in approach temperature. Additionally, this decrease in cost will be at a faster rate than the cost of a comparable water cooler. This is because of the limitation usually imposed on the cooling water return temperature. There is normally no similar restriction on the airside of an air fin. Approach temperatures above 25°F to 40°F generally favor air cooling, however, air cooled exchangers are not normally considered practicable where the approach temperature is less than 15 to 25°F. If a number of services require cooling to an approach less than 25°F it is usually economical to provide air coolers for even these services instead of water cooling equipment if these services are only a small percentage of the total duty. Where it is necessary to cool below a process terminal temperature of 15°F above the design air temperature, consideration should be given to the following alternatives:

2.1.4

1.

Conventional water cooling throughout the range.

2.

Air cooling to an economic approach temperature followed by water cooling to the required final temperature.

3.

Combined evaporative and air-cooling.

Cost Comparison In a location where adequate supplies of cooling water are not available the choice of an air cooler is obvious. In most cases, however, it will be necessary to compare respective costs of air and water cooling before a choice is made. A list of points to be considered in a cost comparison of the two systems is given below. In general, an air cooled exchanger will be more expensive in capital cost than a water cooled exchanger considered without its attendant offsite facilities and less FOSTER WHEELER ENERGY LIMITED 2002

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expensive than a water cooled exchanger with these facilities. Air Cooled Exchanger 1.

Capital cost of exchanger including motors, foundations, supporting structure, electric switchgear and cable, control system.

2.

Shipping

3.

Construction

4.

Power consumed, preferably calculated based on the annual average temperature.

5.

Maintenance

Water Cooled Exchanger

2.2

1.

Capital cost of exchanger including foundations, supporting structure, control system.

2.

Shipping

3.

Construction

4.

Maintenance

5.

Cooling Water Distribution

6.

Proportion of offsite facilities, e.g., cooling tower and pumping costs.

7.

Utility costs, i.e. power, water make-up, and chemicals for treating.

Air Fin Optimization For a given process duty the optimum exchanger in terms of economics is that for which the sum of capital cost of equipment plus running cost of the air fan is a minimum. Thus the calculation to determine the optimum air cooler depends upon the relative costs of power, taken over the payout period, and the exchanger surface specified. The surface for the optimum exchanger will depend upon the correct selection of tube and fin dimensions, number of rows and design air velocity.


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A definite design is normally sought from a specialist vendor but the discussion below indicates the effect of various parameters of interest to the process engineer in preparing preliminary design. 2.2.1

Selection of Tube and Fin Dimensions Tubes - The tube length is usually restricted by the client, who requires a standard size or by installation and maintenance requirements. Occasionally the tube length may be dictated by the width of a pipe rack. Generally, longer tubes result in a lower air fin investment cost. Shorter tubes find use in services, which require a critically lowpressure drop. Tube diameters of 1" are generally specified although process considerations such as pressure drop and velocity limitations may dictate the use of other tube diameters. Fins - The number of fins per inch is decided by the magnitude of the tubeside heat transfer coefficient taken in conjunction with the applied fouling factor. The higher the internal heat transfer coefficient, the greater the number of fins per inch required. This should not exceed eleven since beyond that limit boundary layer effects on the air side reduce the air side coefficient.


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2.2.2



Number of Tube Rows An initial selection of the number of tube rows is made from an examination of the service conditions, the most important factors being the level of cooling and the overall heat transfer rate. These factors determine the relative quantity of air required for a given heat duty. High heat transfer rates and small temperature rise on the airside will tend to result in a small number of rows. Generally, the number of tube rows varies between three and six since with less than three rows, the airflow across the bundle tends to be uneven and increasing the number of tube rows beyond six results in higher fan power costs. However, if there is a restriction of plot area, maximum noise levels or tube side drop, this general rule may no longer be applicable. Final selection of the number of rows is based on the economics of investment cost versus operating costs to obtain a minimum overall cost.

2.2.3

Design Air Velocity Increasing air velocity increases power consumption but lowers the cost of exchanger surface because of better heat transmission. Hence the design air velocity selected is based on an economic balance between exchanger and power. Air velocity is governed by the fan power and the fin density. For services with low coefficients, it is usual to use bare tubes, but for successively higher coefficients fin densities up to 11 fins/inch are normal. The optimum technical solution of a given overall heat transfer problem is reached when the airside heat transfer coefficient multiplied by the surface ratio (finned surface/bare tube surface) equals the tubeside heat transfer coefficient.


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3.0



PROCESS SPECIFICATION A copy of the air-cooled exchanger process specification sheet used in America and England, Figure 1 and 2, is included at the end of this section and relative comments are found below: 3.1

Tubeside Heat Exchanged - Prepare alternative specification sheets where the maximum heat duty does not correspond to the maximum viscosity, pour point or flowrate. Fluid Circulated - This describes the nature of the fluid, e.g. gas, liquid, hydrocarbon, etc. The engineer must also state whether the fluid is fouling and advise the nature and concentration of corrosive compounds, if present. For conditions where heavy fouling is expected and where a very low inside heat transfer coefficient results, it is uneconomic to apply a high ratio have finned surface. In addition, cover plate headers may be required for heavy fouling fluids. Fouling Factors - Only the fouling resistance of the fluid on the inside of the tubes is required. The Standards of Tubular Exchanger Manufacturers Association (TEMA) give fouling resistances for various process fluids. Fluid Flowrates - Flowrates corresponding to the design heat duty are required. If this is not the maximum flowrate, specify this also, as it will probably control the pressure drop through the exchanger. Additionally, where a fluid has a high pour point or high viscosity it is mandatory to specify the minimum flowrate. With an associated low allowable design pressure drop laminar flow and concomitant problems of poor heat transfer and plugging in the tubes may result. Fluid Viscosity - Specify the viscosity of the fluid at two temperatures within the cooling range of the process fluid. If the exchanger has a range of duties, specify also the maximum viscosity case. For fluids with very high viscosity, it may be necessary to increase the allowable design pressure drop to ensure turbulent flow. Pour Points - Specify the pour points of fluids that contain waxy deposits or have high viscosities. The exchanger should be designed so that the tube wall temperature at the cold end under conditions of minimum flow, no fouling, and winter design air temperature is at least 10°F above the pour point. FOSTER WHEELER ENERGY LIMITED 2002

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For services where pour point is critical, vendors design should be given close scrutiny to ensure that design has allowed for process requirements. Steam coils and manual louvres are required for start-up operation in services where fluids have high pour points. A rough approximation for the steam requirement can be obtained for 30 ft. tube length by using 30 lbs/hr of steam per foot of bundle width. However, it is recommended that each case be considered separately taking into account the properties of the process fluid and the minimum air temperature. Additionally, the process fluid outlet temperature may be controlled to prevent sub-cooling. Generally, temperature control is only used where the tube wall outlet temperature is less than 20°F above the pour point at the winter minimum temperature and minimum fluid flowrate. Control is accomplished using automatic louvres, 50% or 100% auto-variable fans and/or air recirculation. The vendor may also consider designing the exchanger with co-current flow and/or bare tubes. Allowable Pressure Drop - Suggested pressure drop for various services are given below. However, care should be taken to ensure that the selected pressure drop results in the most economic overall installation. The allowable pressure drop for product cooling and non-critical services should not control the size of the exchanger, as this may result in the uneconomic design which could be avoided by reconsidering the hydraulics of the process circuit. Special consideration is required for wide temperature range cooling of viscous liquids, low pressure gases or condensation of vapors at very low pressures. In these services, pressure drop is a critical requirement, which greatly influences the size of the heat transfer surface.


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TABLE 1 ALLOWABLE PRESSURE DROP Service Liquid Cooling

Allowable Pressure Drop, psi 10

Note 1

Gas Cooling: Operating pressure 15-50 psig

1

Operating pressure 50-250 psig

3

Operating pressure 250-1500 psig

5

Condensing: (atmospheric pressure and above)

Note 2

Total condensation

0.5 min

Note 3

Partial condensation

2-5

Note 3

Notes: 1.

Not valid for viscous fluids.

2.

For vacuum service the selection of an allowable pressure drop should be from the results of an economic study. Pressure drops are usually in the range of 3-5 mm Hg.

3.

For multi-pass air coolers, high pressure drops assure proper flow distribution. Usually a minimum of 3 passes is considered in a tower condenser. The higher pressure drop will also assure proper distribution at lower than design throughput. Condensing Curve Data - For condensing services state whether condensing curve data is provided or straight line condensation can be assumed. Subcooling Requirements - If it is required to subcool a liquid to a temperature lower than the saturation temperature of the vapor, then the air fin must be supplied with a loop seal to ensure the provision of subcooling surface. The exchanger vendor must advise the depth of the loop seal to ensure flooding of the required tube rows for subcooling. Maintaining the required amount of flooded surface is usually difficult due to the fact that most air coolers have shallow bundles i.e. only 3 or 4 tube rows. FOSTER WHEELER ENERGY LIMITED 2002

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Tubeside Velocity - Specify any limitations on tubeside velocity. To minimize fouling minimum velocities are specified where the process fluid contains solids such as catalyst or where the process fluid is water (for these services the velocity is generally not less than 3 ft/sec.) In addition, maximum velocities are sometimes specified to prevent erosion. Split Headers - Multipass services having inlet process temperatures of up to 350°F - 400°F and a design temperature drop of over 200°F should have horizontally split headers to relieve tube to tube sheet stresses between rows. This special provision should also be made where the outlet process temperature is allowed to reduce during conditions of low ambient air temperature and/or low process flow if this results in a temperature range above 200°F across the section. Freeze Points - Process streams that contain water may freeze due to the ambient air temperature being below the freezing point. The same considerations that apply to high pour point streams apply here and also to materials that crystallize at high temperature. Onstream Cleaning - If special arrangements (e.g. extra nozzles) are required to permit onstream cleaning, they must be specified. Header Vents - In condensers used for total condensation but which may contain non-condensibles, a vent should be located in the outlet header. 3.2

Air Side Site Altitude - This is required to establish the air density, which in turn affects pressure drop across the bundle, fan size, blade pitch angle and driver horsepower. Lowest Winter Temperature - This temperature is required to establish the precautions to be taken when operating with high pour or freeze point fluids. This temperature is also used to determine driver H.P. where it is mandatory for the driver to be capable of operation at the low air design temperature with fan blades set at the pitch suitable for the high air design temperature. The temperature is determined in a similar manner to the summer design air temperature (see below). Design Air Temperature - This determines the approach temperature, which is usually the limiting factor for sizing the exchanger. Choosing a high value may result in an unnecessarily large surface whereas a value, which is too low, may adversely affect operation of the exchanger during peak ambient air conditions. FOSTER WHEELER ENERGY LIMITED 2002

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The cost of the air cooled heat exchanger is directly related to the chosen design air temperature and careful consideration should be given to its selection. The actual figure to be used may be specified by the client. If the client does not specify a design air temperature, he should be consulted with regard to the period of time that reduced capacity can be tolerated as a result of the design air temperature being exceeded. Consideration should be given to selection of services on a critical/noncritical basis. For instance, partial condensation of a tower overheads providing reflux can be considered critical whereas product cooling is obviously non-critical. Different design air temperatures may be selected for the two categories. The common practice for determining the design air temperature is to select a temperature from meteorological data, which is not exceeded for more than a percentage of time throughout the year. This percentage of time varies according to the period over which the meteorological data is assessed and on the location but is usually between !"# and 5% based upon annual hourly readings. It is important to remember that consideration must also be given to the effect of local refinery heat sources on the air temperature. Shell Europe suggest that adding 2°C to the design air temperature is satisfactory. See Appendix A for a statistical method to calculate the design air temperature from meteorological data.


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4.0



MECHANICAL DESIGN FEATURES 4.1

Induced or Forced Draught A forced draught unit pushes air across the tube surface with the fan located below the tube bundle. An induced-draught design has the fan located above the bundle and the air is pulled across the tube surface. The two systems are compared below. The selection of induced or forced draught type is generally made by the vendor who makes a selection based on the advantages/disadvantages of each system related to the process, engineering requirements and economics. However, occasions do arise when the process itself may be a significant factor in this selection. 4.1.1

Induced draught 1.

Induced draught units are less likely to recirculate the hot exhaust air since the exit air velocity is !"# to $"#%times that of the forced draught unit. This low susceptibility to air recirculation is especially important for low temperature approaches and for condensers with low condensing temperature (ammonia, freon, propane).

2.

Better protection of tube bundle against sudden temperature changes due to the weather. Towers operating at ambient temperatures and total condensation may fall into a vacuum condition.

3.

More uniform distribution of air flow across tube bundle since the air velocity approaching the bundle is relatively low.

4.

The unit is more suitable for installation above other mechanical equipment such as pipe racks or shell and tube exchangers.


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4.1.2

4.2


Forced draught 1.

For the same duty the forced draught unit requires less horsepower because it moves air at the lowest available temperature and highest density.

2.

The driver, gears and fans are in the cool air stream. The outlet air temperature of induced draught units is restricted to protect the mechanical equipment in the hot exit air stream.

3.

Provides easier access to the tube bundles and fans.

4.

For the majority of vendors is 5 to 6% cheaper, although induced draft can be more economical in low temperature services where the drivers can be located overhead.

Air Humidification In this system water is sprayed into the incoming air stream to pre-cool the air passing to the exchanger. A relatively small amount of water is used. This water has not normally undergone water treatment, hence effective drift elimination is essential to prevent fin corrosion and the deposition of calcium and magnesium salts onto the fins. This system is most effective in climates with a low humidity but is also used when required process outlet temperatures are so low that air cooling alone is not economical.

4.3

Tube Bundle Orientation Horizontal Type - The tube bundle elements are arranged horizontally. This type is usually preferred because it permits simpler arrangement of pipework and convenient exchanger grouping although the plot area required is larger than with other types. A - Frame - The tube bundle elements form a roof. This type requires about 60% only of the plot area necessary for the horizontal type. The most favorable applications are in vacuum plants, very large units and for condensers having liquid after-coolers mounted beneath the condensing elements. One particularly interesting application is for use as a vacuum steam condenser (turbines) in cold climates as it allows counter-current steam flow and water drainage in the tubes (vacuum pulled from the top). Hence water freezing does not occur. For other applications, reference should be made to manufacturers literature. However, some vendors will not use this type of unit because of the effect of crosswinds. V - Frame - The tube bundle elements form an inverted A. These are used in similar applications to the A-Frame but with an induced draught fan. FOSTER WHEELER ENERGY LIMITED 2002

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Vertical Type - The tube bundle elements are stacked vertically. This orientation is used extensively for small air conditioning type units 4.4

Header Type The types of headers most widely used in air coolers are described below. For a preliminary selection of the header type use can be made of the following recommended pressure limitations for each type. Cover Plate

-

up to 30 kg/cm2

Plugs

-

up to 60 kg/cm2

Manifold

-

over 60 kg/cm2

For more detailed information, reference should be made to FW Eng. Std. 23A1. Plug Type - This type has a closed header with plugs opposite to each tube end. Access for cleaning the tubes is through the plugs. Cover Plate - The tube plugs and plug sheet are replaced with a flanged cover plate. With this type of header it is quicker and simpler to clean the tubes than with the plug type, but it is more expensive and requires thicker plates. Cover plates are used where the fluid has a fouling factor greater than 0.004 but their use is limited to services with low design pressures as this type of construction is susceptible to leaks. Manifold - The inlet and outlet headers are of all welded design with a Utube arrangement for the return bends. This type is used for very high pressure services. Split Headers - When the difference between the inlet and outlet temperature of the process fluid exceeds 200°F, in multipass units, one of the headers should be split to allow for differential expansion between the tube passes. 4.5

Tubes and Fins 4.5.1

Tube and Fin Dimensions Ideally tube length, spacing and pitch and the fin height and spacing would vary with the service. Practically, however, manufacturers find it more economical to limit these parameters to a number of standard configurations. Standard lengths for tubes are 8, 10, 12, 16, 20, 24, 30 and 40 ft., the most common being 30 ft. FOSTER WHEELER ENERGY LIMITED 2002

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Tube bundles are available in widths to 13 ft., lengths to 40 ft. And depths to 8 rows. The most commonly used tube diameter is 1" O.D. which has fins varying in height from 0.5" to 0.625" and in spacing from 3 to 11 per linear inch. There are a few special types such as the rectangular finned GEA elliptical tube designed for certain conditions. For details of these reference should be made to manufacturers' literature. Tube pitch varies between 2.0 to 2.75" triangular to ensure good contact between the air and fins. The factors which determine the selection of a particular configuration are discussed in Section 2.2. 4.5.2

Fin Types The type of fin construction must be suitable for the service conditions taking into account such factors as metal temperature, environmental effect and abnormal operating conditions as well as heat transfer rate. The list below gives the available types of finned tubes with their respective characteristics. Additional information can be obtained from F.W. Eng. Std. 23A1. Tension Wound and Wrapped - The steel or aluminum fin is wrapped under tension on the surface of a bare tube. The heat transfer capacity of this type depends on the small contact surface between fin and tube. If differential expansion occurs between fin and tube this type becomes rapidly inefficient and consequently its use is usually limited to a design temperature of 250°F. It also has poor resistance to atmospheric corrosion because of the large area of exposed bare tube. To increase the contact area between the fins and the bare tube the fin base may be bent to form an "L" or "J" shape, and in this form can be used for design temperature up to 500°F. Mechanically Embedded - The fins are mechanically wrapped under tension and embedded in a grove spirally cut into the outside surface of the tube. This type is suitable for high temperature service (up to 750°F design temperature) because it has good distribution of contact surface and low bond resistance. Welded, Brazed and Soldered - The fins are attached to the tubes by brazing or soldering instead of grooving or tension winding. The use of the tubes is limited by the softening temperature of the brazing compound or solder (i.e. virtually no limit on design temperature). FOSTER WHEELER ENERGY LIMITED 2002

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Welded fin tubes are costly and are used only in special cases for severe duties. Bimetallic - In this type of tube an aluminum liner from which fins have been formed by extrusion is bonded to an inner tube. Good thermal efficiency is ensured by a large contact area but at high tube wall temperatures bond resistance becomes acceptable. Integral - This is the best type of finned tube but also the most expensive. The completely integral fin ensures that any mechanical or heat transfer problem which may exist with a bonded fin is eliminated and, therefore, this type of finned tube may be used for the most severe temperature and pressure services, e.g. aluminum suitable for use with wall temperatures up to 250-280°F.


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4.6



Fans and Drivers 4.6.1

Fans Fan Pitch - Fans are generally of two types; manually adjustable where the pitch angle can be adjusted when the fan is at rest and auto-variable in which the pitch angle can be automatically controlled during operation. In either case the pitch normally has a range of 30°of positive and 10-15°of negative angle. The negative pitch can be used to suppress natural convection flow across the tube bundle and prevent overcooling during cold ambient conditions. Fan Layout - To ensure even distribution of air flow across the bundle face, each air fin should be provided with a minimum of two fans. In addition, the fan area should not be less than 40% of the plot area. This rule can be relaxed for induced draught designs because the velocity of the air entering the tube bundles on the suction side is much lower than with the forced draught type; the value for induced draught designs is 30%. Fan Blades - Generally four blades are employed in modern air coolers unless there is a noise restriction or severe service requirement (high air flow rate and high static pressure) in which case six or eight bladed fans are used.

4.6.2

Drivers and Speed Reducers Electric motors are the most widely used drivers and they should be rated to operate at the specified minimum ambient air temperature. The most common types of speed reducer are V-belts or right angle gear drives. V-belts and pulleys are cheaper in initial capital cost and are often preferred from a maintenance view point as failures are virtually restricted to belt breakages and belts can be replaced quickly. For motor HP's above 30 only gear drives are suitable, but because of their high initial cost they should only be used when an economic balance dictates the necessity to save exchanger surface area by increasing power consumption. When the fan diameter is small (up to 5 ft.) the speed reducer is eliminated and the motor is connected directly to the fan.


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4.7



Materials of Construction 4.7.1

Header, Cover, Plugs, and Gaskets Headers, header covers and plugs are normally fabricated from carbon steel. Where a corrosion problem exists these items are fabricated from alloy steel or carbon steel with a large corrosion allowance. Specifications for gasket materials are given in FW Eng. Std. 23A1. Plug gaskets are generally solid metal but metal jacketed asbestos filled type with the metal compatible with the header material can also be used. Cover plate gaskets are compressed asbestos unless the vendor or client has overriding requirements.

4.7.2

Tubes and Fins Prime tubes are normally carbon steel either seamless or electric resistance welded. Seamless tubes only are used in hydrogen service or where corrosive attack on the seam of a welded tube may occur. Alloys and non-ferrous tubes are sometimes specified to overcome corrosion problems, e.g. aluminum brass for a crude overheads exchanger. Aluminum is the most economic and commonly used fin material. FW Eng. Std. 23A1 specifies temperature limitations for fin material.

4.7.3

Fans The most commonly used materials are plastics and cast aluminum. Cast aluminum is generally acceptable but can be subject to corrosion in certain atmospheres. This can be overcome by use of corrosion resistant paint. Plastic or laminated fiberglass blades are now accepted industrially and are corrosion resistant, although they should be limited to 200°F maximum air temperature. Steel blades are sometimes used on larger fans but tend to be expensive. Carbon steel blades require a protective coating. Wooden blades with or without a plastic coating are no longer accepted by FW.


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4.7.4



Corrosion Allowance A minimum corrosion allowance of 1/8" should be specified on all internal carbon steel or low alloy surfaces exposed to the process stream except the tubes unless otherwise specified by a client. Higher allowances may be used for particularly corrosive services. It should be noted that the tube wall temperature in air fins is marginally higher than in water cooled shell and tube exchangers and corrosion rates may be different.

4.7.5

Design Pressure and Temperature Design Pressure - The location of the air fin usually dictates the method which is used to determine its design pressure. If the air fin is located in a pump discharge circuit and can be shut in against pump discharge, the design pressure is normally set at the shut-off pressure of the pump calculated by the method given in Section 400 of the Process Design Standards. Alternatively the airfin may be part of a system such as a distillation tower and condenser which is protected by one relief valve. In this case, the design pressure is set at the maximum operating pressure in the system plus 10% or 25 psig whichever is the greater. Design Temperature - Specified as the maximum process fluid temperature plus 50°F.

4.8

Distribution Piping It is usually impractical to have a completely symmetrical piping arrangement at the air fin inlet or outlet. To minimize the effect on the flow distribution of non-symmetry, the pressure drop in the distribution piping should be small in comparison with the total overall system pressure loss. In general, the preferred piping system is as shown in type A&B with inlet and outlet headers running the full length of the air fin header box with direct connections from the header to each pass. Type A (Co-Current)


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The characteristics of this type are: Simple arrangement Low Pressure Drop (1) Poor flow distribution and hence unbalanced heat transfer. Type B (Counter Current)

The characteristics of this type are: Simple arrangement Low pressure drop (1) Poor flow distribution but better than Type A.


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This is the preferred type for liquid flow. (See F. W. Francaise report "Straight Manifold Design for Balanced Flow" for additional information.) Type C

The characteristics of this type are: More complex structure High pressure drop Good flow distribution In multipass two phase service the numerous bends and T's promote fluid separation. For additional information on flow distribution see Chemical Engineering June 17, 1968, pages 210-213. One known refinery has solved it's two phase mal-distribution problems by adding notched stand pipes to the individual inlets in the inlet header. Apparently the notched stand pipes distribute the condensed liquid to all inlets allowing the vapor to readily flow through each coil pass. (1).

The design of piping distribution systems should be reviewed very carefully, particularly in large units operating at low pressure and where piping frictional losses are to be minimized for vapor services.


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5.0



OPERATION AND CONTROLS 5.1

Multiple Fans Where a unit is equipped with multiple fans a simple control may be achieved by taking selected fans out of service. This has an added advantage due to savings in power and still provides the required degree of cooling if the air temperature is lower than the design. Air fin exchanger publications (e.g., Hudson Engineering) state that with an induced draught design up to 40% of the cooling duty may be achieved by natural convection when the fans are not in operation; the value for forced draught designs is 15 to 20%.

5.2

Process Bypasses Product outlet temperature control can be provided by bypassing some of the process fluid round the air cooler through a conventional control valve in the bypass. In the case of an air fin with a low &P, e.g. a condenser, a butterfly control valve should be used. The method is readily applicable to multi-service units and is accurate but wasteful of power. Residue coolers and other high viscosity or high pour point streams cannot use this method since there is a serious risk of the cooler plugging due to overcooling of the process stream in the cooler.

5.3

Adjustable Louvers Louvers may be used to control the product outlet temperature by restricting the flow of air through the tube bundle. Manual louvers are used for adjustment of product temperatures which may vary due to process or air temperature changes. Automatic louvers may be used to control the product outlet temperature of a critical service which is arranged in a common unit with other services under one or more common fans. Generally, the initial cost for automatic louvers is higher than for AV fans and they have the further disadvantage that the fans consume close to full power at all times. One advantage of louvers is that they provide some protection to the tube bundle during inclement weather.


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5.4



Variable Pitch Fans Variable pitch fans are commonly used to provide accurate control of fluid outlet temperature under conditions of varying heat load and varying ambient air temperatures. The pitch of the fans may be adjusted either manually or automatically. The use of auto-positioners for the fan blades also gives a saving in horsepower as the fans move only the required amount of air. Auto variable pitch fans cannot be used where a number of heat exchangers for different services share common fans as auto fans operated from the fluid outlet temperature of one cooler would affect the operation of the other coolers. On very large single service units employing a number of fans, accurate control and economical operation may be achieved by using a proportion of auto-variable fans (say 50%) with the remainder being manually adjustable. Hysteresis curves shall be required from manufacturers when using automatic pitch adjustment. This is particularly important when the air fins are tower condensers.

5.5

Air Recirculation Where it is necessary to condense or cool fluids which solidify above the ambient air temperature, recirculation of warm air is used to prevent solidification in the cooler. Part of the warm air leaving the tube bundles is passed through the fan again and mixed with the correct proportion of fresh air to give a constant cooling air temperature. The control of the mixing ratio between warm and fresh air is by means of louvres or dampers which can be operated automatically or manually. For start-up or part load operation in winter, heating elements may also be built- in. Air recirculation systems are very expensive and with very large duties such as residue coolers in refineries there is an economic limit to the application of this design. As an alternative the engineer should consider cooling the fluid liable to solidification on the shellside of a shell and tube exchanger using tempered water from an auxiliary circuit. The water in the auxiliary circuit is then cooled using a normal fin tube cooler. This alternative should in fact be considered every time a viscous material is to be cooled in an air fin. It is particularly important for large services with laminar flow conditions in the air fin when the combination of shell and tube + airfin + tempered water is cheaper and saves plot areas.

5.6

Variable Speed Motors Product outlet temperature control can be achieved by varying the speed of the fans. This can be done by variable gear boxes, hydraulic drives or FOSTER WHEELER ENERGY LIMITED 2002

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variable speed motors, but these methods are usually expensive in terms of capital and running costs. Variable speed motors are generally used for small fans and low horsepower requirements only. Coarse regulation, to save power, can be achieved by the use of two-speed motors.


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6.0



PLOT PLAN Airfin Location Airfin exchangers are supported from grade on their own structure or on top of some other structure, a frequent location being above an elevated pipe rack; the space beneath the structure is used for other equipment. In either case certain precautions are necessary. Exchangers should not be located over or near equipment giving off large quantities of heat and the amount of equipment under the exchangers should be minimized and arranged so as not to impede air flow or promote air recirculation. In addition, location of air fins at different levels should be avoided as this can also promote warm air recirculation. If airfins are unavoidable near to heat sources, then the design air temperature should be raised accordingly. Also to be avoided are parallel banks of airfins which are so close together as to impede the flow of cold air. Airfin Type Roof-type exchangers are sometimes used for very large units as they require only about 60% of the floor space of horizontal types. Induced draught horizontal types have some advantages over forced draught horizontal types in plot plan considerations. For a given column height there is more space beneath the induced draught type which can be utilized for the installation of other equipment. Additionally, when overhead drives are used, the induced draught type is more suitable for mounting over existing equipment and pipe racks. Tube Bundle Size For units with a small heat transfer surface, consideration should be given to reducing the tube length and increasing the bundle width to obtain a more compact unit. Combination of Services The plot area required can be minimized by grouping together under common fans services which are non-critical, i.e. services such as product coolers which do not require accurate control of their outlet temperatures. Units which are too small to have independent fans are also combined. Tower condensers should not be combined as accurate control may be required to maintain required tower conditions. Fan Mounting To ensure satisfactory air distribution, fans should be mounted at least one-half the fan diameter (minimum distance is 6'-6") above grade for forced draught designs FOSTER WHEELER ENERGY LIMITED 2002

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and above the tube bundle for induced draught designs. Column Top Condensers GEA, a German vendor with many foreign licensees, appears to have good experience in the location of airfin condensers at the top of medium sized fractionating towers and this system may be considered particularly for partial condensers.


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7.0



NOISE Maximum allowable operating noise levels for air cooled exchangers are specified in the Foster Wheeler Engineering Standards 23A1. The noise emanating from airfin coolers is essentially axial fan noise. The power of the noise generated is a function of the fan speed and blade angle and it can be reduced by using a slower speed fan. The fan may be equipped with a greater number of blades to move the same quantity of air. The use of intake and discharge silencers on the air flow has not generally been used except on small plants near residential areas or in air conditioning systems due to the considerable cost involved.


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8.0



ESTIMATION OF EQUIPMENT SIZE AND UTILITY CONSUMPTION This section provides the process engineer with a method for determining the approximate tube surface, operating power and plan area of the air cooled heat exchanger. The major factors affecting the calculation for an economic design such as air velocity, tube and fin dimensions and number of tube rows are discussed in detail in Section 2.2. The method given below should not be used as a rigorous solution to the thermal design for the optimum exchanger which depends upon the correct selection of the above parameters and the relative costs of power and tube surface. 8.1

Calculation Procedure 1.

Overall Coefficient - Assume an overall heat transfer coefficient (Ud) from Table 2 or from Figure 3 if the process fluid is all liquid and has a known viscosity. Care should be taken with highly viscous fluids which have very low overall coefficients (Ud = 5 to 20). In these cases a small change in the coefficient can result in a large change to the required surface area.

2.

Log Mean Temperature Difference - Assume a hot air outlet temperature (Tao) and calculate the LMTD. Apply the correction factor (Ft) from Figure 4 using the curves applicable to the selected number of passes. The minimum value for (Ft) should be 0.8. For a good approximation of the air temperature rise the following equation can be used: &tair = Ufinned x 0.12 (Average fluid temp. - Air inlet temp.) Select the number of tube passes from the following guide: Gas cooling and condensers

1 pass

Cooling of medium/heavy hydrocarbons

2 passes

Cooling of light hydrocarbons

4 passes


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3.


Surface Area - Calculate the bare tube surface area from the equation:

A=

Q U D x LMTD x Ft

A = outside bare tube surface (sq. ft.) Q = heat exchanged (Btu/hr) 4.

Number of Tube Rows - Evaluate the ratio:

Ti - Tai and from Figure 5 obtain the number of rows (NR). U Ti = process fluid inlet temperature (°F) Tai = design air inlet temperature (°F) 5.

Bundle Face Area - Calculate the bundle face area (FA) from the bare tube surface area using the ratios given in Table 3.

6.

Air Outlet Temperature - Select a face velocity (FV) from Table 4 corresponding to the calculated number of tube rows and calculate the air outlet temperature from the equation Tao = Tai +

Q FA x FV x 1.08

(The density and specific heat of air at the standard conditions, 70°F and 29.92 inches of Hg, are incorporated in the constant 1.08.) Compare the calculated value of Tao with the assumed value from (2). If the values do not check within 5% assume a new value for Tao and repeat the calculations (2) to (5). 7.

Air Flow Rate - Calculate the actual volumetric air flow (ACFM) from the equation ACFM = FA x FV x ' Where ' =

1 Air Specific Gravity

(See Figure 6)


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For forced draught units ' is measured at the air inlet temperature, Tai, and at the air outlet temperature, Tao, for induced draught units. 8.

Airside Static Pressure Drop - Estimate the static pressure drop (&P) from Figure 7, using the face velocity from (6) and correct for the number of tube rows and actual air density to obtain the static pressure drop (SP) using the equation: SP = &P x

NR x '% (inches of water) 3

' is determined at the arithmetic mean air temperature. 9.

10.

Fan Selection - From the tube length and face area calculated in (5) calculate the unit width and estimate the fan dimensions using the following criteria: (a)

Provide a minimum of two fans per unit.

(b)

The total area circumscribed by the fan blades must be at least 40% of the face area for forced draught designs and 30% for induced draught designs.

(c)

Use a standard size fan. See Figure 8.

Total Pressure Drop - Obtain the velocity pressure drop (VP) from Figure 8 and calculate the total pressure drop (&P) from the equation

&P = SP + VP (inches of water) 11.

Power Consumption - Absorbed horsepower per fan is estimated from the equation HP (fan) =

ACFM x &P 3900 x Number of fans

Size the fan driver for 120% of absorbed fan horse power to allow for climatic and process changes.


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8.2


Sample Problem Process Duty: Service

Kerosene Cooler

Heat exchanged

17,700,000 Btu/hr

Kerosene inlet temperature

250°F

Kerosene outlet temperature

130°F

Design air inlet temperature

86°F

Tube dimension:O.D. (bare)

1"

Length Airfin type

30'-0" Forced draught

Calculation: 1.

Use UD = 90 Btu/hr-ft2-°F

2.

Assume Tao = 145°F

(Table 2)

LMTD = 70.1°F (uncorrected)

P=

145 - 86 = 0.36 250 - 86

R=

250 - 130 = 2.035 145 - 86

Assuming 4 tube passes Ft = 1.0 (Figure 4) Corrected LMTD = 1 x 70.1°F

3.

Bare tube surface area, A =

17,700,000 ft 2 90 x 70.1

=

2,805 ft 2 FOSTER WHEELER ENERGY LIMITED 2002

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4.

Ti - Tai UD



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=

250 - 86 = 1.82 90

Number of tube row, NR = 5.3 (Figure 5) Use 5 tube rows 5.

6.

Bundle face area =

2,805 ft 2 6.32

=

444 ft 2

(Table 3)

Face velocity = 565 ft/min Tao = 86 +

(Table 4)

17,700,000 (F 444 x 565 x 1.08

= 151(F This checks within 5% of the assumed value, 145°F 7.

Air flow rate at Standard Conditions

8.

Static pressure drop, &P = 0.27 ins. H2O (Figure 7) Corrected static &P = 0.27 x

5 x 1.0888 in. H2O 3

(' at mean air temp. of 115°F from Figure 6) = 0.488 ins. H2O 9.

Fan Selection: Unit width =

444 = 14.8 ft. 30

Required fan area 0.4 x 444 = 178 ft2 Consider 2 fans, 1/2

fan diameter

=

. 178 x 4 + , ) - 2/ *

=

10.6 ft.

Use 2 x 11 ft. diameter fans. FOSTER WHEELER ENERGY LIMITED 2002

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10.



Total Pressure Drop Velocity pressure drop, VP = 0.143 ins. H2O (Figure 8) Total pressure drop, &P = 0.488 + 0.143 ins. H2O = 0.631 ins. H2O

11.

Horsepower per fan absorbed = 255,000 x 0.631 = 20.6 BHP 3900 x 2 H. P. of each driver = 1.2 x 20.6 = 24.7 Therefore, use 25 HP motor for each driver.


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TABLE 2 TYPICAL OVERALL HEAT TRANSFER COEFFICIENTS

All the overall heat transfer coefficients (UD) for air coolers listed below are based on the total bare outside surface of the tubes. Liquid Cooling (Also see Figure 3) Service LPG

UD (Btu/hr ft 2 °F) 110 - 125

Propane Butane Process Water

105 - 120

Light Naphtha

90 - 110

Gasoline Pentane Platformate Heavy Naphtha

85 - 95

Kerosene Light diesel Medium Hydrocarbons

65 - 75

Atmospheric gas oil Heavy diesel Atmospheric bottom pumparound Heavy Hydrocarbons

40 -50

Reduced crude Light vacuum gas oil Fuel Oil

20 - 30

Residuum

10 - 20

Tar

5 - 10


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Gas Cooling Service Air or flue gas at 50 psig

UD (Btu/hr ft2o F) 10

(&P = 1 psi) Hydrocarbon gases at 15 - 50 psig (&P = 1 psi) Hydrocarbon gases at 50 - 250 psig

30 - 40 50 - 60

(&P = 3 psi) Hydrocarbon gases at 250 -1,500 psig

70 - 90

Ammonia Reactor Stream

80 -90

Condensing Service

UD (Btu/hr ft2°F)

Steam (0 - 20 psig)

130 – 140

Amine Reactivator

90 - 100

Ammonia

100 - 120

Light Hydrocarbons

80 - 95

Ethane Propane Butane Light Naphtha

70 -80

Reactor Effluent

60 - 80

Platformers Hydroformers Heavy Naphtha

60 -70

Still Overhead

60-70

(Light naphtha, steam and non-condensible gas) Freon 12

60-80


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TABLE 3 BUNDLE FACE AREA RATIOS

Tube Rows

3

4

5

6

7

8

9

10

11

12

* sq. ft. surface per sq.ft. Face Area

3.80

5.04

6.32

7.60

8.84

10.08

11.36

12.64

13.92

15.20

* Based on bare tube surface Table taken from Hudson Engineering Corporation Technical Data Booklet.

TABLE 4 TYPICAL FACE VELOCITIES Tube Rows Face Velocity, ft/min

3

4

5

6

7

8

9

10

11

12

630

595

565

540

510

490

465

445

425

405

Table taken from Hudson Engineering Corporation Technical Data Booklet.


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APPENDIX A CALCULATION OF DESIGN AIR TEMPERATURE FROM METEOROLOGICAL DATA The following statistical procedure can be used for calculating the design air temperature from recorded meteorological data. Assume that the temperatures recorded follow the normal Gaussian distribution. Let 0 = the standard deviation of all temperatures. 0 = 1/6 (Absolute maximum temperature - absolute minimum temperature) The Gaussian distribution shown below can be represented as: Q1x 2 4

1 2/

e

3x2 2

A

Min.

Mean

T

Max.

Temp No.of SD

Let T be the required temperature which is not exceeded for r% of the year. Probability integral t = T mean - T 0 Area shaded (A) = 0.5 -

r 100

The value of t is found from standard tables.


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Example Calculate the temperature exceeded for 1% of the year from given meteorological data. Assume absolute Min. temp. = 27°F (over one year) Assume absolute Max. temp. = 114°F (over one year)

0=

Then the S.D.

114 - 27 = 14.5 6

T mean (based on average daily max. and min.) = 58.5°F Shaded area (A) = 0.5 -

1 = 0.49 100

From tables t = -2.33 Substitute in

t=

T mean - T

0

- 2.33 =

58.5 - T 14.5

T = 92°F The following table indicates the effect on the temperature T by a change in the percentage of the year when T is exceeded. r%

Area A.

t (from tables)

T, °F

1

0.49

2.33

92.3

2

0.48

2.05

88.2

3

0.47

1.88

85.7

4

0.46

1.75

83.9

5

0.45

1.64

82.3

Normal practice is to use the temperature which is not exceeded for 5% of the year.


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FWC air cooled exchanger.pdf

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