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IPS-E-PR-755(1)
This Standard is the property of Iranian Ministry of Petroleum. All rights are reserved to the owner. Neither whole nor any part of this document may be disclosed to any third party, reproduced, stored in any retrieval system or transmitted in any form or by any means without the prior written consent of the Iranian Ministry of Petroleum.
ENGINEERING STANDARD
FOR
PROCESS DESIGN OF FANS AND BLOWERS
FIRST EDITION
JULY 2014
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FOREWORD
The Iranian Petroleum Standards (IPS) reflect the views of the Iranian Ministry of Petroleum and are intended for use in the oil and gas production facilities, oil refineries, chemical and petrochemical plants, gas handling and processing installations and other such facilities.
IPS is based on internationally acceptable standards and includes selections from the items stipulated in the referenced standards. They are also supplemented by additional requirements and/or modifications based on the experience acquired by the Iranian Petroleum Industry and the local market availability. The options which are not specified in the text of the standards are itemized in data sheet/s, so that, the user can select his appropriate preferences therein
The IPS standards are therefore expected to be sufficiently flexible so that the users can adapt these standards to their requirements. However, they may not cover every requirement of each project. For such cases, an addendum to IPS Standard shall be prepared by the user which elaborates the particular requirements of the user. This addendum together with the relevant IPS shall form the job specification for the specific project or work.
The IPS is reviewed and up-dated approximately every five years. Each standards are subject to amendment or withdrawal, if required, thus the latest edition of IPS shall be applicable
The users of IPS are therefore requested to send their views and comments, including any addendum prepared for particular cases to the following address. These comments and recommendations will be reviewed by the relevant technical committee and in case of approval will be incorporated in the next revision of the standard.
Standards and Research department
No.17, Street14, North kheradmand
Karimkhan Avenue, Tehran, Iran.
Postal Code- 1585886851
Tel: 021-88810459-60 & 021-66153055
Fax: 021-88810462
Email: [email protected]
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IPS-E-PR-755(1)
GENERAL DEFINITIONS:
Throughout this Standard the following definitions shall apply.
COMPANY:
Refers to one of the related and/or affiliated companies of the Iranian Ministry of Petroleum such as National Iranian Oil Company, National Iranian Gas Company, National Petrochemical Company and National Iranian Oil Refinery And Distribution Company.
PURCHASER:
Means the “Company” where this standard is a part of direct purchaser order by the “Company”, and the “Contractor” where this Standard is a part of contract documents.
VENDOR AND SUPPLIER:
Refers to firm or person who will supply and/or fabricate the equipment or material.
CONTRACTOR:
Refers to the persons, firm or company whose tender has been accepted by the company.
EXECUTOR:
Executor is the party which carries out all or part of construction and/or commissioning for the project.
INSPECTOR:
The Inspector referred to in this Standard is a person/persons or a body appointed in writing by the company for the inspection of fabrication and installation work.
SHALL:
Is used where a provision is mandatory.
SHOULD:
Is used where a provision is advisory only.
WILL:
Is normally used in connection with the action by the “Company” rather than by a contractor, supplier or vendor.
MAY:
Is used where a provision is completely discretionary.
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CONTENTS: PAGE No.
0. INTRODUCTION ............................................................................................................................. 1 1. SCOPE ............................................................................................................................................ 2 2. REFERENCES ................................................................................................................................ 2 3. DEFINITIONS AND TERMINOLOGY ............................................................................................. 2 4. SYMBOLS AND ABBREVIATIONS ............................................................................................... 4 5. UNITS .............................................................................................................................................. 5 6. GENERAL ....................................................................................................................................... 5
6.1 Fan Identification .................................................................................................................... 5 6.2 Pressure Limit of Application ................................................................................................ 5 6.3 Types of Fan ............................................................................................................................ 5 6.4 Performance ............................................................................................................................ 6 6.5 High Temperature Service ..................................................................................................... 6
7. DESIGN CRITERIA ......................................................................................................................... 6 7.1 Selection Parameters of Process Fans................................................................................. 6 7.2 Operational Characteristics and Performance .................................................................... 6 7.3 Fan Laws .................................................................................................................................. 8 7.4 Performance Calculations ...................................................................................................... 8 7.5 Fan Control .............................................................................................................................. 8 7.6 Fan Systems .......................................................................................................................... 10 7.7 Fan Selection Procedure ...................................................................................................... 10 7.8 Process Data Sheet ............................................................................................................... 13
APPENDICES: APPENDIX A TYPICAL PERFORMANCE CURVES OF THE COMMONLY USED FAN TYPES 15 APPENDIX B .................................................................................................................................. 16 APPENDIX C FAN SELECTION GUIDE CHART .......................................................................... 18 APPENDIX D PERFORMANCE CALCULATION AND FAN SELECTION ................................... 21 APPENDIX E TYPES OF FANS & BLOWER ................................................................................ 26
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0. INTRODUCTION
"Pressure Reducing/Increasing Machineries and/or Equipment" are broad and contain variable subjects of paramount importance. Therefore, a group of process engineering standards are prepared to cover the subject.
This group includes the following standards:
STANDARD CODE STANDARD TITLE
IPS-E-PR-745 "Engineering Standard for Process Design of Vacuum Equipment (Vacuum Pumps & Steam Jet Ejectors)"
IPS-E-PR-750 "Engineering Standard for Process Design of Compressors"
IPS-E-PR-755 "Engineering Standard for Process Design of Fans and Blowers"
This Standard covers:
"Process Design of Fans and Blowers"
Basic process engineering calculation related to fans and blowers such as interpretation of performance curves, efficiency, power requirement, preparation of data sheets, etc. are presented.
As a general rule, all gas compressing equipment which produce less than 35 kPa pressure at discharge, with an atmospheric (or slightly sub-atmospheric) suction pressure, fall into the category of "Fans and Blowers".
The terms "fan" and "blower" are used interchangeably by the industry. The American Society of Mechanical Engineers (ASME) uses specific ratio, to separate fans (specific ratio up to 1.11), blowers (specific ratio 1.11-1.20), and compressors (specific ratio above 1.20).
In this standard, some of the subjects are adapted from the following specifications and handbooks:
- “Applied Process Design”,vol.3, 2&3th Edition, by Ernest Ludwig
- Design Practice by EXXON Engineering, “section XI-G(FANS)”, 2001
- “Energy Conservation Standards Rulemaking Framework for Commercial and Industrial Fans and Blowers”, 2013, by the U.S. Department of Energy (DOE).
- “Power Smart guide book”, prepared by Ivor da Cunha P.Eng., Terry Strack P.Eng., and Saul Stricker P.Eng. of LeapFrog Energy Technologies Inc, issued by CEATI International Inc., 2008.
- “Fan & Pump II”, by Dr. Sam C M Hui, Department of Mechanical Engineering The University of Hong Kong E-mail: [email protected]
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1. SCOPE
This Engineering Standard Specification covers the minimum requirements, basic reference data and necessary formulas for process calculations and proper selection of Fan and Blower to be used in the OGP industries.
It contains basic reference information, data and formulas necessary for fan selection as mentioned above.
Note 1:
This standard specification is reviewed and updated by the relevant technical committee on Jul. 1998. The approved modifications by T.C. were sent to IPS users as amendment No. 1 by circular No. 26 on Jul. 1998. These modifications are included in the present issue of IPS.
Note 2:
This is a revised version of this standard, which is issued as revision (1)-2014. Revision (0)-1993 of the said standard specification is withdrawn.
2. REFERENCES
Throughout this Standard the following dated and undated standards/codes are referred to. These referenced documents shall, to the extent specified herein, form a part of this standard. For dated references, the edition cited applies. The applicability of changes in dated references that occur after the cited date shall be mutually agreed upon by the Company and the Vendor. For undated references, the latest edition of the referenced documents (including any supplements and amendments) applies.
API (AMERICAN PETROLEUM INSTITUTE)
API Standard 673 "Centrifugal Fans for Petroleum Chemical and Gas Industry Services"
ANSI/API Standard 661 "Air-Cooled Heat Exchangers for General Refinery Service"
IPS (IRANIAN PETROLEUM STANDARDS)
IPS-E-GN-100 "Engineering Standard for Units"
IPS-E-PR-250 "Engineering Standard for Performance Guarantee"
UIPS-M-PM-230U "Material and Equipment Standard for Special Purpose Centrifugal Fans"
3. DEFINITIONS AND TERMINOLOGY
TERMS USED IN THIS STANDARD ARE DEFINED AS FOLLOWS:
3.1 FAN IMPELLER
The assembly of the fan wheel and the hub(s).
3.2 Fan Plane
A flow area perpendicular to the flow of gas at the specified reference plane; that is, inlet flange or outlet flange.
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3.3 Fan Rated Point
Defined as (1) the highest speed necessary to meet any specified operating condition and (2) the rated capacity required by fan designs to meet all operating points. (The Vendor shall select this capacity point to the best encompass specified operating conditions within the scope of the expected performance curve.).
3.4 Maximum Continuous Speed (rotations per minute)
The speed at least equal to the product of 105 percent and the highest speed required by any of the specified operating conditions.
3.5 Normal Operating Point
The point at which usual operation is expected and optimum efficiency is desired. Unless otherwise specified, fan performance shall be guaranteed at the normal operating point.
3.6 Specific Ratio
Total pressure at the fan outlet divided by the total pressure at the fan inlet, measured at peak fan total efficiency.
3.7 System Component Resistances
The total system pressure loss or resistance is the sum of the resistances of individual component parts, such as ducts, enlargements, contractions, filters, etc.
3.8 Tip Speed Also called peripheral velocity, equals the circumference of the fan wheel time the RPM of the fan and is expressed in m/s (ft/min).
3.9 The Total Pressure (Ptf) of a Fan
The rise of pressure from fan inlet to fan outlet as measured by two impact tubes, one in the fan inlet duct and one in the fan discharge duct, corrected for friction to the fan inlet and outlet respectively.
Where no inlet duct is used, total pressure on the inlet side is zero.
3.10 The Velocity Pressure (Pv) of a Fan
The pressure corresponding to the average velocity determination from the volume of air flow at the fan outlet area.
3.11 The Static Pressure (Ps) of the Fan
The total pressure (Ptf) diminished by the fan velocity pressure (Pv).
3.12 Standard Air Density
1.2014 kg/m³ (0.075 lb/ft³) corresponding to an ambient temperature of 20°C (70°F) pressure of 1.013 Bara (14.7 psia).
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3.13 The Unit of Pressure
The mm. of water column of density of 997.423 kg per cubic meter and/or Pa (1 mm H2O conventional= 9.80665 pascals).
3.14 The Volume Handled by a Fan
The number of cubic meters of air per hour expressed at fan outlet conditions.
3.15 The Power output of a Fan
The power output of a fan is the power supplied to the air stream and is termed air power. It’s expressed in kilowatts and is based on fan volume and fan total pressure.
3.16 The power input to a fan
Expressed in kilowatts and is the measured kilowatt delivered to the fan shaft.
3.17 The Mechanical Efficiency of a Fan
The ratio of power output to the power input.
3.18 The Static Efficiency of a Fan
The mechanical efficiency multiplied by the ratio of the static pressure to the total pressure or es = et × Ps /Pt.
3.19 The Fan outlet Area
The inside area of the fan outlet.
3.20 The Fan inlet Area
The inside area of the fan inlet collar.
3.21 Evase
A diffuser or a diverging discharge transition piece.
4. SYMBOLS AND ABBREVIATIONS
BkW = Brake (shaft) kilowatt of Fan in (kilowatts, kW);
D = Wheel Diameter, in (m);
d = Relative Density, (dimensionless);
es = Static Efficiency, in (fractions);
et = Mechanical (Total) Efficiency in (fractions);
F1 = Temperature Correction Factor, in (kg/m3);
F2 = Altitude Correction Factor, in (kg/m3);
FkW = Fan Power, in (kilowatts);
GkW = Gas (Air) kilowatt of Fan, in (kilowatts, kW);
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K = Ratio of Specific Heats, Cp/Cv, (dimensionless);
P1 = Fan Inlet Pressure, in [mm H2O (abs.)], or in [Pa (abs.)];
Ps = Static Pressure of Fan, in [mm H2O (abs.)], or in [Pa (abs.)];
Ps2 = Fan Outlet Static Pressure, in [mm H2O (abs.)], or in [Pa (abs.)];
Pt = Total Pressure in (mm H2O), or in (Pa);
Ptf = Fan Total Pressure in (mm H2O), or (Pa);
Pv = Velocity Pressure of Fan, in (mm H2O), or (Pa);
pv2 = Fan Outlet Velocity Pressure, in [mm H2O (abs.)], or [Pa(abs.)];
r/min = Rotational Speed, in (rotations per minute);
T1 = Gas Temperature at Fan Inlet, in (K);
V1 = Fan inlet Rate, in (m3/h);
Vm = Gas Velocity, in (m/s);
Vp = Peripheral Velocity, in (m/s);
t = Temperature Rise, in (K or degree °C);
ρ (rho)= Density (Mass Density), in (kg/m3);
π (pi) = Constant, equal to 3.1416;
Subscripts:
t = Based on total pressure;
s = Based on static pressure;
1 = At inlet conditions;
2 = At outlet conditions.
5. UNITS
This standard is based on international system of units (SI), as per IPS-E-GN-100 except where otherwise specified.
6. GENERAL
6.1 Fan Identification
Fans are rather generally identified as machines with relatively low pressure rises which move air or gases or vapors by means of rotating blades or impellers and change the rotating mechanical energy into pressure or work on the gas or vapor. The result of this work on the fluid will be in the form of pressure energy or velocity energy, or some combination of both.
6.2 Pressure Limit of Application
Fans for all services handling air or gas, on process duties (excluding those for direct cooling or ventilating) and which develop less than 35 kPa (0.35 bar) from atmospheric pressure, shall conform to API Std. No. 673 (see 2.1), and Iranian Petroleum Standards IPS-M-PM-230, for " Forced Draft Fans for Boilers and Process Services."
6.3 Types of Fan
For types of fan refer to Appendix E of this Standard.
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6.4 Performance
6.4.1 Fan performance shall be based on the static pressure rise across the fan inlet and outlet flanges. In specifying required operating conditions on the data sheet, the purchaser is responsible for accounting for inlet velocity pressure. The fan vendor is responsible for including the pressure losses attributed to all items within his scope to obtain the required static pressure rise. 6.4.2 Unless otherwise specified, fans shall have turndown capability to 60% or less of rated flow. For parallel operations, fan performance shall have a continuously rising pressure characteristic (pressure versus flow plot) from the rated capacity to surge. Performance curves, corrected for the specified gas at the specified conditions, shall be based on performance tests of actual or prototype equipment, including eaves, if any, and inlet box(es).
6.4.3 Fan performance shall be guaranteed to meet all operating conditions specified on the data sheet and shall be within the tolerances listed, at the normal operating conditions.
a) For variable-speed fans, the static pressure and capacity shall be guaranteed with the understanding that the power shall not exceed +3%. These tolerances shall not be additive.
b) For constant-speed fans the specified capacity shall be guaranteed with the understanding that the static pressure shall be within +5% and -0% of that specified; the power shall not exceed stated power by more than +3%. These tolerances shall not be additive.
6.4.4 Fan performance shall be guaranteed in accordance with IPS Standard IPS-E-PR-250, "Performance Guarantee".
6.5 High Temperature Service
High temperature service for fans and auxiliaries is defined as a service for air and/or gas with a fan inlet temperature greater than 200°C. Fans, in this service will be indicated as "High temperature fans".
Centrifugal fan models are available for gas temperatures to 1000°F (540°C) for induced combustion draft and hot gas recirculation services. Air circulation impellers (“cooling wheels”) are placed on the shaft between the casing and bearing to shield the bearings and driver coupling form the high temperature. Oil lubricated bearings with circulating oil systems are used. The speeds which fan designers employ are reduced below speeds used for ambient temperature services, typically to 96% of maximum at 450°F (230°C) and 75% at 800°F (425°C). Radial blading is normally used so that the required head can be obtained with minimum tip speed. MACHINERY SPECIALIST should be consulted when specific applications are considered.
7. DESIGN CRITERIA
7.1 Selection Parameters of Process Fans
This section of standard is intended to cover information necessary to determine the approximate power requirement and other selection parameters of process fans and to furnish proper data for evaluation of manufacturer’s proposals and/or preparation process data sheet.
7.2 Operational Characteristics and Performance
7.2.1 Centrifugal fans
The three types of centrifugal fan blades (radial, backward and forward) give three characteristic performances. Figures A.1 to A.3 of Appendix "A" present the typical performance curves for radial, backward and forward bladed fans respectively. Exact performance for a given fan can only be obtained on test.
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a) Radial blade
This type of blade is usually used for handling suspended materials, abrasive dust collecting and exhausting of pumps from dirty, greasy or acid environment.
The rather sharply rising static pressure curve of the radial blade centrifugal fan allows for small changes in volume as the resistance of the system changes considerably.
A Fair running static efficiency is 50-70 percent for both the straight radial blade and radial tip blade.
b) Backward blade
This type of blade is well suited to stream line conditions and is used extensively on ventilating, air conditioning and clean and dirty process gas streams. The outstanding and important characteristic is the non-overloading power (kilowatts). It eliminates the need for oversized motors or other drivers. The usual operating static efficiency range for the regular blade is 65-80 percent and for the streamlined design is 80-92 percent.
c) Forward blade
This type is usually shallow and operates at slow speed for a given capacity and usually has low outlet velocity.
Its operating characteristics are poor for many applications, since the power rises sharply with a decrease in static pressure once the peak pressure for the fan has been reached. The operating static efficiency range is 55-75 percent.
7.2.2 Axial flow fans
The typical performance of the axial flow fan is represented in figure A.4 of Appendix "A".
The power characteristic is none overloading. The usual pressure range of application is 0-76 mm water (0-745 Pascal) static pressure. The vane axial and tube axial can be selected for higher outlet velocities than the centrifugal (10.2 - 20.3 m/s).
The axial fans should be connected to ducts by tapered cone connection. The peak efficiency range of the tube axial is 30-50 percent and for the vane axial is 40-65 percent.
7.2.3 Propeller fans
These fans usually operate with no piping or duct work on either side, and move air or gas from one large open area to another. Pressures are usually very low and volumes depend upon size, blade pitch, number of blades and speed. Wall and ceiling mounted exhausters (like “attic fans") and ceiling fans are examples of the simple, low power types. Cooling tower and air cooled heat exchanger fans are examples of the engineered, high power type.
Efficiency is typically 20% lower than for the tube axial type if the partial housing (or orifice) is designed for smooth flow transition. Efficiency can be as low as 10 to 20% for crudely designed orifices. Wide blades of the household fan type tend to be quiet but low in efficiency. Narrow blades are more efficient but noisier.
For good air flow distribution in air-cooled heat exchangers, the fan diameter is selected so that the fan area is greater than 40% of the total base area of the tube bank.
Static efficiency of long bladed propeller fans is typically 30 to 35%. However, in the last few years major fan blade manufacturers have developed and offer low and super low noise designs that in addition to the low noise advantage (see below) can achieve efficiencies in the 50 -60 % range. Tip speeds for air cooled heat exchanger fans are limited to 200 ft/sec (60 m/sec) by GP 6-4-1 (API 661) to limit noise generation with low noise designs typically in the 60 ft/sec (18 m/sec) to 100 ft/sec (30 m/sec) range. Average air velocity across the fan area is in the range of 1200 to 2500 ft/min. (6 to 13 m/sec).
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7.3 Fan Laws
The performance of a fan is usually obtained from a manufacturer’s specific curve. Fan Performance can be estimated for points other than the normal flow and speed point by use of the fan laws. For a given fan these laws are: (1) Volume varies directly with speed (2) pressure rise varies with the square of speed and directly as the density, and (3) power varies as the cube of speed and directly as the density.
FAN LAWS
Fig. 1
Fan laws apply to blowers, exhausters, centrifugal and axial flow fans. The relations are satisfactory for engineering calculations as long as the pressure rise is not greater than 7 kPa.
Theoretically a 100 mm H2O (0.98 kPa) pressure rise affects air density to cause a one percent deviation. Where greater accuracy is required, the familiar adiabatic power relations are used.
These laws are applicable only for geometrically similar fans and to the same point of rating on the performance curve.
7.4 Performance Calculations
For calculating of fan performance, a number of parameters must be determined including total system pressure, tip speed of fan, power of fan, air velocity, temperature rise, and ….
A typical calculation of performance is provided in Appendix D.
7.5 Fan Control
The fan control parameter may be inlet pressure, discharge pressure, flow, or some combination thereof. This may be accomplished by suction throttling (by a damper or variable inlet guide vanes), or discharge blow off (when a constant speed drive is used). Variable fan speed may also be used. The control system may be mechanical, pneumatic, hydraulic, electrical, or any combination thereof. The system may be manual or it may be automatic with a manual over-ride.
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7.5.1 For constant speed drives, the control signal shall actuate an operator that positions an inlet and/or outlet damper or adjustable inlet guide vanes. (Also called radial inlet damper, variable inlet vanes, vortex dampers.)
7.5.2 For variable speed drives, the control signal shall act to adjust the set point of the driver’s speed control system. Unless otherwise specified, the control and operating speed range shall be from the maximum continuous speed to 95% of the minimum speed required for any specified operating case or 70% of maximum continuous speed, whichever is lower.
7.5.3 The full range of the specified control signal will correspond to the required operating range of the driven equipment. Unless otherwise specified, the maximum control signal shall correspond to the maximum continuous speed or the maximum flow.
7.5.4 Facilities shall be provided to automatically open or close (as specified) the dampers/guide vanes on loss of control signal and to automatically lock or brake the dampers or vanes in last position on loss of motive force (air supply, electric power, etc.).
Note: This is a specific system consideration and the associated controls should be arranged to avoid creating hazardous or other undesirable conditions.
7.5.5 The fan vendor shall furnish and locate the operators, actuator linkages, and operating shafts for remote control of the dampers/guide vanes. Operator output shall be adequate for the complete range of damper or guide vane positions. The proposed location of operators, linkages, and shafts shall be reviewed with the purchaser for consideration of maintenance access and safety.
7.5.6 External position indicators shall be provided for all dampers/inlet guide vanes.
7.5.7 Fan volume is controlled by the following methods:
a) Variable Speed Drive This type of control can be accomplished by turbines, DC motors, variable speed motors or slip-ring motors. With changing speed of the driver the fan output capacity and pressure can be varied. For capacity reductions below 50 percent, an outlet damper is usually added to the system. b) Outlet Damper with Constant Fan Speed The system resistance is varied with this damper. The volume of gas delivered from the fan is changed as a function of the movement of the damper. It is low in initial cost and simple to operate, but does require more power than other methods of control. c) Variable Inlet Vane with Constant Fan Speed The angle and/or extent of closure of the inlet vanes controls the volume of gas admitted to the inlet of the wheel. The inlet vane control is more expensive than the outlet dampers but this can usually be justified by lower kilowatt costs, especially on large power installations. d) Fluid Drive This method allows fan speed to be adjusted 20-100 percent with corresponding volume changes. A constant speed motor is used, see Fig.2, note that curve F of this figure is the actual power input to the fan shaft. The hydraulic of fluid drive has about 3 percent in losses, so its power input at 100 percent load is actually about 103 percent to allow for this. Curves B and C are for variable vane inlet dampening and Curve A is for outlet dampening of a backward blade fan. Curve E shows an outlet damper with multiple step speed slip-ring motor. This has outlet damper for final control from 89-100 percent. From this graph a reasonably accurate selection can be made of the control features to consider for most installation conditions.
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COMPARISON OF EFFICIENCIES OF FIVE PRINCIPAL METHODS OF CONTROLLING FAN
OUTPUT
Fig. 2
7.6 Fan Systems
An operating fan is a part of some system. Regardless of the system, the fan cannot be selected until the flow and resistance characteristics have been analyzed. Fan selection for the system is based on the static pressure for a given volume of gas flowing. Since most fans operate at relatively low pressures the effect of uncertainty or error in resistance calculations can have a large percentage effect on Kilowatt and operational characteristics. Since it is essentially impossible to determine exact figures for the system resistances. It is to add 10-20 percent to the calculated static pressure as a safety factor. The fan discharge pressure necessary to overcome the system resistance is composed of friction loss plus losses due to changes in velocity (accelerations and decelerations) in the duct and connections of the system.
In order to analyze the total system resistance and its relation to fan performance, refer to Appendix C.
7.7 Fan Selection Procedure
The following steps should be followed in fan selection.
7.7.1 Specifying the fan type
Informations presented in sections 6.3 and 7.2 may help in selecting the suitable fan for the process. Fan type curves should also be studied in order to recognize the effects of changes in system resistance on the fan performance and the volume and pressure changes caused by variations on speed. Recommendations of manufacturers are of particular importance in this stage.
Fig. C.1 of Appendix "C" may be used in fan type selection.
7.7.2 Specifying the inlet volume
The volume of a fan should be determined by (1) the process material balance plus reasonable extra (about 20 percent) plus volume for control at possible future requirement, (2) generous capacity for purging, and (3) process area ventilation composed of fume hoods, heat dissipation and normal comfort ventilation. Table 1 gives suggested air changes for area ventilation, but not air conditioning.
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TABLE 1 - AVERAGE AIR CHANGES REQUIRED FOR GOOD VENTILATION MINUTES PER CHANGE MINUTES PER CHANGE
Assembly halls……………………………………… 2-10 Auditoriums………………………………………….. 2 -10 Bakeries……………………………………………… 2 - 3 Banks…………………………………………………. 3-10 Barns………………………………………………… 10 - 20 Bars………………………………………………….. 2 - 5 Beauty parlors……………………………………… 2 - 5 Boiler rooms……………………………………….. 1 - 5 Bowling alleys……………………………………… 2 -10 Churches……………………………………………. 5 -15 Clubs………………………………………………… 2 -10 Dairies………………………………………………. 2 - 5 Dance halls…………………………………………. 2 -10 Dining rooms………………………………………. 3 -10 Dry cleaners……………………………………….. 1 - 5 Engine rooms………………………………………. 1 - 3 Factories……………………………………………. 2 - 5 Forge shops………………………………………. 2 - 5 Foundries………………………………………….. 1 - 5 Garages……………………………………………. 2 -10
Generator rooms…………………………………. 2 - 5 Gymnasiums……………………………………….. 2 -10 Kitchens – hospital………………………………… 2 - 5 Kitchens – resident……………………………….. 2- 5 Kitchens -restaurant……………………………… 1- 3 Laboratories……………………………………….. 1- 5 Laundries…………………………………………… 1 - 3 Markets……………………………………………... 2-10 Offices ………………………………………….. 2-10 Packing houses…………………………………… 2- 5 Plating rooms……………………………………… 1- 5 Pool rooms…………………………………………. 2 - 5 Projection rooms…………………………………. 1- 3 Recreation rooms………………………………… 2-10 Residences………………………………………… 2- 5 Sales rooms……………………………………….. 2-10 Theaters……………………………………………. 2- 8 Toilets………………………………………………. 2- 5 Transformer rooms………………………………. 1- 5 Warehouses……………………………………….. 2-10
7.7.3 System resistance
The system resistance must be calculated in the usual manner and at the actual operating conditions of the fan.
Corrections are then applied to convert this condition to "standard" for use in reading the rating tables.
7.7.4 Manufacturers multirating tables
The multirating tables (rating tables) of fan manufacturers are convenient for selecting any of the many types of fans. Usually m3 /h values can be found close enough to requirements to be acceptable. Direct interpolation in the table for volume, r/min and BkW is acceptable for narrow ranges, otherwise the fan laws must be used.
Performance tables are based on dry air at 21°C at sea level (barometric pressure 760 mm mercury) with a density of 1.2 kg/m3.
When the fans are required to handle gases at other conditions at the inlet, corrections must be made for temperature, altitude and air or gas density.
7.7.4.1 Performance at conditions other than above:
1) Calculate actual density of gas (or air) under operating conditions.
For Air, Fig. 3 is convenient to use:
Read temperature correction factor, F1
Read altitude correction Factor, F2
Actual density = 2.1
))(( 21 FF , Kg /m3 (Eq.13)
at operating conditions
For gases other than air the density must be calculated since the curves of Fig. 3 are for an air density of 1.2 kg/m3. If the gas density at 21 °C and 101.325 kPa is close to that of air under these conditions, then the curves could be used for convenience. Otherwise, calculate the actual gas density by the gas laws.
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2) Calculate the equivalent static pressure:
Equivalent static pressure
density Actual2.1
(Eq.14)
3) From manufacturers’ rating tables for air or gas, using the required m3/h at inlet operating conditions and the equivalent static pressure calculated in (2), read the r/min (rpm) and BkW. Interpolate if necessary.
4) The r/min is the correct value for the actual operating conditions.
5) The BkW must be corrected for density:
(Eq.15)
6) The correct performance at the actual operating conditions will be:
m3/h as set at inlet conditions
Static pressure as set at inlet conditions, mm of water
Temperature as set at inlet conditions, °C.
r/min as read from manufacturers’ tables.
BkW as corrected by (5) above. Examples for fan selection are presented in Appendix "D".
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* Note: Conditions of air as defined in para. 7.7.4.
CORRECTION FACTORS FOR EFFECTS OF ALTITUDE AND TEMPERATURE ON AIR*
Fig. 3
7.8 Process Data Sheet
The following data should be prepared by process engineer as the basic informations to be filled in data (specification) sheet;
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a) Gas characteristics
Air
Process gas; Name:
Composition:
Mol. Mass:
b) Operating conditions
Capacity, (dm3/s or kg/h)
Relative density, (Air = 1)
Barometric pressure, [m bar or mm H2O (abs.)]
Elevation above see level, meters (m)
Relative humidity, (%)
Inlet temperature (°C)
Inlet pressure [m bar or mm H2O (abs.)]
Outlet pressure [m bar or mm H2O (abs.)]
c) Control
Outlet Dampers, (std.), (stream flow)
Variable inlet vanes, (manual), (auto)
Variable speed drive
Slip ring motor
Variable pitch blades
Other
d) Fan Type
Centrifugal
Vane axial
Tube axial
Propeller
Other
See Appendix F for typical data sheet.
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APPENDICES
APPENDIX A
TYPICAL PERFORMANCE CURVES OF THE COMMONLY USED FAN TYPES
CENTRIFUGAL
BACKWARD-INCLINED BLADING
Fig. A.1
CENTRIFUGAL
RADIAL BLADING
Fig. A.2
CENTRIFUGAL
FORWARD-CURVED BLADING
Fig. A.3
AXIAL FLOW FAN
Fig. A.4
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APPENDIX B
TABLE B.1 - TABLE OF FAN LAWS Fan Law
Number
Ratio of
Variables * Normally Constant **
Ratio × Ratio
Ratio
a 1 b c
m3/h press.
Kw
=
size3 × (r/min) size2 × (r/min)2 size5 × (r/min)3
×
1 d d
a 2 b c
m3/h r/min kW
=
size2 × press.1/2 1/size × press.1/2 size2 × press.3/2
×
1/d1/2 1/d1/2 1/d1/2
a 3 b c
r/min press. KW
=
1/size3 × (m3/h) 1/size4 × (m3/h)2 1/size4 × (m3 /h)3
×
1 d d
a 4 b c
m3/h press. r/min
=
size4/3 × kW1/3 1/size4/3 × kW2/3 1/size5/3 × kW1/3
×
1/d1/3 1/d1/3 1/d1/3
a 5 b c
size r/min kW
=
(m3/h)1/2 × 1/press.1/4 1/(m3/h)1/2 × press.3/4 (m3/h) × press.
×
d1/4 1/d1/4
1 a 6 b c
size press kW
=
(m3/h)1/3 × 1/(r/min)1/3 (m3/h)2/3 × (r/min)4/3 (m3/h)5/3 × (r/min)4/3
×
1 d d
a 7 b c
size m3/h kW
=
press.1/2 × 1/(r/min) press.3/2 × 1/(r/min)2 press.5/2 × 1/(r/min)2
×
1/d1/2 1/d3/2 1/d3/2
a 8 b c
size r/min press.
=
1/kW1/4 × (m3/h)3/4 kW3/4 × 1/(m3/h)5/4 kW × 1/(m3/h)
×
d1/4 1/d3/4
1 a 9 b c
size r/min m3/h
=
kW1/2 × 1/press.3/4 1/kW1/2 × press.5/4 kW × press.
×
d1/4
1/d3/4 1
a 10 b c
size m3/h press
=
kW1/5 × 1/(r/min)3/5 kW3/5 × 1/(r/min)4/5 kW2/5 × (r/min)4/5
×
1/d1/5 1/d3/5 d3/5
(to be continued)
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TABLE B.1 (continued) FOR CONSTANT AMOUNT BY MASS (SIZE CONSTANT). WHEN DENSITY OF AIR VARIES. 11 m3/h, r/min AND PRESSURE VARY INVERSELY AS THE DENSITY, THAT IS, 1BINVERSELY AS THE BAROMETRIC PRESSURE AND DIRECTLY AS THE ABSOLUTE TEMPERATURE. 12 KILOWATT VARIES INVERSELY AS THE SQUARE OF THE DENSITY THAT IS INVERSELY AS THE SQUARE OF THE BAROMETRIC PRESSURE AND DIRECTLY AS THE SQUARE OF THE ABSOLUTE TEMPERATURE. FOR CONSTANT AMOUNT BY MASS (SIZE CONSTANT). WHEN BOTH TEMPERATURE AND PRESSURE VARY. 13 m3/h AND r/min VARY AS √ PRESSURE × ABSOLUTE TEMP. KILOWATT VARIES AS √ PRESSURE3 × ABSOLUTE TEMP.
* Either one of these variables may be used independently in which case the other must remain constant. Then the desired data will vary as power of the variable indicated. These laws are written as if both varied at the same time. In this case the desired data will vary as the powers of the variables indicated.
** In the last column the effect of relative density (d) is shown.
This is to apply as an extra variable when needed, otherwise density is supposed to be constant.
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APPENDIX C
FAN SELECTION GUIDE CHART
TYPICAL FAN SELECTION GUIDE CHART, BASED ON PRESSURE
RISE VERSUS AIR FLOW
Fig. C.1
Notes on Fig. C.1:
1) This chart shows the range of centrifugal and axial machines, and is based on catalogue ratings. The standard centrifugal ventilation fans operate approximately 560 mm of water. Beyond this, heavy-duty centrifugal fans, with higher compression ratios, may be made to specifications. The only area where no fan is available is above 2500 mm H2o at extremely low flows.
2) To use this graph: (1) calculate the actual m3/h at inlet conditions (at fan flange), and the total pressure rise-in mm of water-from the inlet to the discharge fan-flange; (2) locate the m3/h value on the chart; if the region where the point falls can be served by more than one type of fan (axial versus centrifugal, or different types of axial or centrifugal), decide on the type of fan by making economic and engineering evaluations.
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TABLE C1- COMPARISON BETWEEN VARIOUS TYPES OF FANS
Backward, airfoil
centrifugal fan
Forward-curved
centrifugal fan Vane-axial Propeller fan
Fan total pressure Δptf Flow rate Fan power input Fan modulation Fan total efficiency Sound power level Airflow direction Volume and weight Initial cost Applications
Higher Δpt All flow rates Nonoverloading Inlet vanes AC inverter 0.7 to 0.86 Lower, higher Lw At low frequencies 90° turn Greater Higher Large HVAC&R system
Comparatively lower Δpt Larger flow rate Overloading Inlet vanes AC inverter 0.6 to 0.75 Medium, higher Lw At low frequencies 90° turn Less Medium Lower pressure, Small HVAC&R systems
Higher Δpt AII flow rates Nonoverloading Controllable pitch AC inverter 0.7 to 0.88 Medium, difference of Lw values is small at various Hz Parallel to axle Greater Higher Large HVAC&R systems
Low Δpt Larger flow rate Nonoverloading 0.45 to 0.6 Higher, higher Lw At high frequencies Parallel to axle Medium volume and lower weight Low Low-pressure, high-volume flow Exhaust systems
(Source: Wang, S. K., 2001. Handbook of Air Conditioning and Refrigeration)
C3
Without defining what comprises the system resistance, but representing it by Curve A–A, this system is to flow 13,000 cfm of air at 1.1 in. static pressure. A fan has been selected that operates at 600 rpm and is represented by its static pressure Curve C–C. The intersection of these two curves, Point 1, is the only point of operation for the system. It is not the exact point required for this system, as Point 1 represents a flow of 13,600 cfm at 1.19 in. static pressure and the requirement of 3.87 hp. To obtain the exact conditions of the problem, the fan discharge may be dampered. The damper resistance will have to be equivalent to the fan pressure at 13,000 cfm (1.19 in. static pressure) minus the required 1.1 in. static pressure, which is equal to 0.09 in. The horsepower from Curve D–D will be 3.83 hp.
The fan will now have a system resistance Curve A–B and operate at Point 2. As an alternate approach to securing a system balance at the point required, the motor speed can be changed by a suitable means. If the fan speed is reduced by 13,000/13,600, the new speed should be (0.889)(600) = 573 rpm. A new fan Curve E–E will go through the desired point conditions. The new horsepower for this operation will be (3.87)(573/600)3 = 3.35 hp.
Note that the second scheme requires less running horsepower but does require adjustment to a new fan speed to produce Curve E–E. They will now operate at Point 3 without a damper. If the resistance of the system had included a back-pressure of 1/2 -inch water as shown by system Curve F–F, the volume of 13,000 cfm would be reached at some speed greater than 600 rpm and located at Point 4 on the static pressure Curve G–G.
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SYSTEM RESISTANCES
Fig. C.2
TYPICAL GRADIENT CURVES OF PRESSURE THROUGH A FAN SYSTEM. (USED BY
PERMISSION: THE HOWDEN FAN COMPANY.)
Fig. C.3
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APPENDIX D
PERFORMANCE CALCULATION AND FAN SELECTION
D.1 Performance Calculation
D.1.1 Pressure
a) Total System Pressure
The sum of the static and velocity pressure is the "Total System pressure", Pt.
Pt = Ps + Pv (Eq. 1)
Where:
Ps is static pressure
Pv is velocity pressure
The fan total pressure Ptf, (see 3.9) is measured as the increase in total pressure given to a gas passing through a fan. It is a measure of the total energy increase per unit volume imparted to the flowing gas by the fan. The static pressure is the fan total pressure less than the fan outlet velocity pressure, (see 3.11).
b) Velocity Pressure (see 3.10)
2( )
19.608m
VVP ρ
= mm of water (Eq. 2)
or: 2( )
2m
VVP ρ
= Pa (Eq. 3)
Where:
ρ is gas density
Vm is gas velocity,
Velocity pressure is indicated by a differential reading of an impact tube facing the direction of air flow in the fan outlet. It is a measure of the kinetic energy per unit volume of gas, existing at the fan outlet.
D.1.2 Peripheral velocity or tip speed
The peripheral velocity of the fan wheel or impeller is expressed as:
60p
D NV π= (Eq. 4)
Where:
Vp is peripheral velocity
D is wheel diameter
N is speed in RPM
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D.1.3 Power
a) Fan kilowatt based on total pressure:
(Eq. 5)
Where:
(FKW) is fan kilowatt V1 is inlet rate Pt is total pressure
b) Fan kilowatt based on static pressure output:
(Eq. 6)
Where:
Ps is static pressure c) Gas kilowatt (Air kilowatt) output:
(Eq. 7) d) Shaft or brake kilowatt (input), based on direct current motor:
BKW = (Amp.) (Volts) (Motor efficiency) x 10-3 (Eq. 8) e) Shaft or brake kilowatt (input), based on alternating current (3-phase) motor:
BKW = 3(Amp.) (Volts) (Motor efficiency) (Power factor) ×10-3 (Eq. 9)
D.1.4 Efficiency
a) Mechanical (total) efficiency:
(Eq. 10)
b) Static efficiency
(Eq. 11)
D.1.5 Temperature rise
The temperature rise as the gas passes through a fan is:
(Eq. 12)
Where:
Δt = Temperature rise, °F. T1 = Air or gas temperature at fan inlet, °Rankine Ps2 = Fan outlet static pressure, inches water absolute; or other absolute units
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P1 = Atmospheric pressure or fan inlet pressure (if not atmospheric), inches water absolute; or other absolute units. Pv = Fans outlet velocity pressure, inches water absolute; or other absolute units. k = Ratio of specific heats, Cp /Cv. es = Fan static efficiency, fraction
D.2 Fan Selection (Example)
A system requires 10300 m3/h of air at 205°C against a 52 mm H2O static pressure. The installation is at elevation 430 meters.
a) Determine: what type of fan blading should be used. A backward-curved blade fan will be selected for this installation because the following are not known: (1) the accuracy with which the system characteristic of 52 mm of water at 10300 m3/h was determined, (2) the type of process control to be used, and (3) the possible system variation.
A backward-curved blade will take care of the above unknowns. It will have:
1) High efficiencies. It is a blade that offers flexibility by inherently providing high efficiencies over a wide range. It has its highest efficiency near its maximum kW-power. This gives flexibility above and below the design point.
2) Non-overloading characteristics. A backward-curve blade will allow close "motoring" without fear of overloading in the event of process upsets.
3) Steep static pressure curves. It offers a wide range of static pressures with a small change in capacity.
b) Determine: can the 10300 m3/h at 205°C and against 52 mm of water, (2.04-inch) static pressure be used to select a fan from the manufacturers’ tables?
The manufacturers’ tables are prepared in accordance with the industry standard set up by the Air Moving and Conditioning Association (AMCA). These tables are based on standard air.
Operating conditions other than these must be corrected before going into the table.
Since this is an air system the density correction chart Fig. 3 is used:
1) Actual density of air at operating condition:
From chart, read at 205°C:
F1 = 0.74
Also read from the lower curve at 430 m altitude:
F2 = 1.15
Actual density = (0.74)(1.15)/1.2 = 0.709 kg/m3 (from Eq. 13)
Air density ratio = 0.709/1.2 = 0.59
2) Equivalent static pressure (at standard conditions):
= 52×709.02.1
= 88 mm of water
3) Select fan from manufacturers’ performance table, at 88 mm of water and 10300 m3/h. Suppose that the nearest acceptable unit in the manufacturers’ table has the following characteristics:
Inlet volume = 10330 m3/h
Speed = 2064 r/min at 88.9 mm H2O (3.5 inches).
BKW = 3.862 kW (5.18 hp) for standard air.
4) Actual r/min = 2064 (would be almost the same for actual conditions).
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5) Actual BKW of fan = 3.862 (0.709/1.2) = 2.282 kW.
6) Performance at 430 m elevation and 205°C inlet air temperature will be: (approximately, this can be improved by applying fan laws to data read from table).
Inlet volume = 10330 m3/h
r/min = 2064 (±)
BKW = 2.282 kW
Determine: could any other fan be used in this application? The next larger or smaller fan size should be examined. Other manufacturers could possibly give a different size that might be more efficient. The final selection should be based on an analysis of several different manufacturers’ fans.
Determine: what is the tip speed of this fan?
Wheel diameter = 460 mm (18.5 inches).
use Eq. 4 (see 7.4.2)
Determine: what is the outlet velocity of the fan.
When quietness of operation is important, the outlet velocity should be in the range of 6 to 11 m/s.
The low outlet velocity corresponds to low outlet velocity pressure, and as this latter factor directly influences power consumption. The velocity should be kept to a minimum, particularly when the static pressure is low. However, it should be pointed out that very low outlet velocities (less than 5 m/s) are not really desirable because they produce no advantages, not even quietness. Actual decibel ratings can be attained from the manufacturer, and these are the best indication of actual noise level to be expected.
Example for fan selection using a process gas
A fan is to handle 84000 m3/h (at suction conditions) of a process gas at a suction condition of 49°C and 93 kPa (abs.) and is to discharge at 64 mm, H2O. The gas density at these suction conditions is 1.36 kg/m3.
Since manufacturers’ tables are based on standard 1.2 kg/m3 air density, this density difference must be recognized.
According to Fan Law No. 6, if the r/min (speed) and m3/h (capacity) are constant, the pressure and power both vary directly as the relative density.
1) Equivalent static pressure
2) Read the fan characteristics (nearest) from the manufacturers’ table. Suppose the following information is extracted:
Wheel dia. = 1.524 m, (60 inches)
Max. peak BKW = 139.23( / min)
1000r
Outlet area = 0.0133 m2
m3/h (listed) = 84400
Outlet velocity = 12.2 m/s
Static pressure = 57.15 mm H2O
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BKW = 19.48 kW (26.1 hp)
r/min = 521
Since the capacity of the fan is slightly over the requirements, interpolation and correction of speed and BKW will not be made (0.47 percent over).
3) Correct r/min = 521 (not interpolating).
4) Actual m3/h = 84400.
5) Actual BKW KW=19.481.361.2
× =22.08 kW
6) Max. peak BKW=139.2 3521( )1000
× =19.68 kW (for Air)
Max. peak BKW=19.681.36( )1.2
× =22.31 kW (for gas)
This indicates that the selection is operating at near peak condition.
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APPENDIX E TYPES OF FANS & BLOWER
How to Recognize Fan & Blower Types
• The air is pressurized by the aerodynamic lift generated by the fan blades, much like a propeller or an airplane wing.
• Although they can sometimes be used interchangeably with centrifugal fans, axial fans are commonly used in ‘‘clean air,’’ low-pressure, high-volume applications.
• Axial fans have less rotating mass and are more compact than centrifugal fans of comparable capacity.
• Additionally, axial fans tend to require higher rotational speeds and are somewhat noisier than in-line centrifugal fans of similar capacity.
a) Centrifugal Fans
Centrifugal fans are rugged, are capable of generating high pressures with high efficiencies and can be manufactured to accommodate harsh operating conditions. These are the most commonly used types of industrial fans.
Centrifugal fans have several types of blade shapes, including:
• Backward-inclined curved blade;
• Backward-inclined, airfoil blade;
• Backward-inclined, flat blade;
• Forward curved;
• Radial-blade; and
• Radial-tip.
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Fig. E.1:Typical characteristics and applications of centrifugal fans
Type Characteristics Applications
Forward curved
(See Fig. E.2)
- blades curve in the direction of rotation
- compared to other types, have low efficiency (between 55 and 65 percent)
- small size relative to other fan types
- low speed, does not require high-strength design
- relatively quiet
- limited to clean service applications
- fan output is difficult to adjust accurately (note how the fan curve is somewhat horizontal), and these fans are not used where airflow must be closely controlled.
- the power curve increases steadily with airflow as the back-pressure drops
- applications that require low to medium air volumes at low pressure
- well suited for residential
heating, ventilation and air conditioning (HVAC) applications
- careful driver selection is required to avoid overloading the fan motor.
- the dip in the performance
curve represents a potential stall region that can create operating problems at low airflow rates.
Radial-blade, (See Fig. E.3)
- suitable for low to medium airflow rates at high pressures
- capable of handling high-particulate
airstreams, including dust, wood chips and metal scrap because the flat blade shape limits material build-up
- blades can be coated with protective
compounds to add resistance to erosion and corrosion
- even in stall situations where vibrations
can be a problem, large clearances between the blades also allow this fan to operate safely and quietly
- many rugged industrial applications
- “workhorse” of industry
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Type Characteristics Applications
Radial-tip,
(See Fig. E.3)
- fills the gap between clean-air fans and the more rugged radial-blade fans. This type is more efficient than forward curved and radial blade fans because of reduced turbulence resulting from the low angle of attack between the blades and the incoming air.
- well-suited for use with airstreams that have small particulates at moderate concentrations and airstreams with high moisture content
- efficiencies up to 75 percent
- used in airborne solids - handling services because they have large running clearances.
Backward-inclined, flat
(See Fig. E.4)
- flat blades are inclined in the direction opposite to the rotation
- considered more robust than other types
- the low angle of contact with the airstream facilitates the accumulation of deposits on the fan blades
- performance drops off at high airflow rates
- suitable for forced-draft service. (Fan is exposed to the relatively clean airstream on the upstream side of the process.)
- unsuitable for airstreams with airborne particulates.
- “safe” choice because of its non-overloading motor characteristic
- often selected when system behavior at high airflow rates is uncertain
Backward-inclined, curved
(See Fig. E.4)
- curved blades inclined away from the direction of rotation
- more efficient than flat blades
- low angle of contact with the airstream promotes the accumulation of deposits on the fan blades
- performance drops off at high airflow rates
- suitable for forced-draft service. (Fan is exposed to the relatively clean airstream on the upstream side of the process)
- because of its non-overloading motor characteristic, this fan type is often selected when system behavior at high airflow rates is uncertain
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Type Characteristics Applications
Backward-inclined, airfoil
(See Fig. E.4)
- airfoil blades tilt away from the direction of rotation
- most efficient with thin blades (~85%),
but most unstable because of stall
- low angle of impingement with the airstream promotes the accumulation of deposits on the fan blades as well as erosion.
- performance drops off at high airflow rates
- suitable for forced-draft service. (Fan is exposed to the relatively clean airstream on the upstream side of the process.)
- because of its non-overloading motor characteristic, this fan type is often selected when system behavior at high airflow rates is uncertain
Radial blade centrifugal fans are capable of serving widely varying operating conditions, which can be a significant advantage in industry.
FORWARD-CURVED CENTRIFUGAL FAN
Fig. E.2
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RADIAL-BLADE AND RADIAL-TIP CENTRIFUGAL FANS
Fig. E.3
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BACKWARD-INCLINED CENTRIFUGAL FANS
Fig. E.4
Figure E.5 shows a comparative diagram of various fan characteristics including the relationship between pressure and flow.
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COMPARATIVE FAN DESIGNS AT AN EQUAL POWER CONSUMPTION
Fig. E.5
b. Axial Fans
Axial airflow fans have a number of advantages over other types including:
• Compactness;
• Light weight;
• Low cost;
• Direct-drive units operating near the synchronous speed of the induction motor; and
• Belt-drive units offering flexibility in fan speed selection.
Usual applications for axial fans are:
• Exhausting contaminated air or supplying fresh air;
• Unidirectional or reversible air-flow applications;
• Exhaust applications where airborne particulate size is small, such as dust streams, smoke and steam.
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Disadvantages:
• Axial fans have an undesirable characteristic that can cause problems in situations where the air flow must vary considerably; these fans have a stall region in the lower airflow range that makes them unsuitable for systems operating under widely varying air flow conditions.
• There are anti-stall devices available that can be installed to alter the airflow patterns around the fan blades and virtually eliminate the problem of stall.
• The problem of stall can be avoided by selecting a fan type with a stable fan operation over the entire range of airflow and pressure.
• To achieve the same airflow capacity as centrifugal fans, axial fans must rotate at a higher speed. For this reason, axial fans are generally noisier than comparable centrifugal fans.
• Access to the motor is restricted by the location of the blades and supports.
There are three types of axial fans; their characteristics and common applications are described in more detail in Figure 6.
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Type Characteristics Applications
Propeller Fans
(see Fig. E.7)
- develop high airflow rates and low pressures
- not suitable for extensive ductwork
- relatively low efficiencies
- inexpensive
- comparatively noisy
- power requirements of propeller fans decrease with increasing airflow
- maximum efficiency at lowest delivery pressure
- often used in rooftop ventilation applications
Tubeaxial Fans
(see Fig. E.8)
- achieve higher pressures and better operating efficiencies than propeller fans
- applied in medium-pressure, high air flow rate applications
- the airflow downstream of the fan is uneven, with a large rotational component
- generates moderate airflow noise
- because of low rotating mass, they can quickly accelerate to rated speed
- because of the high operating speeds of 2-, 4-, and 6-pole
induction motors, most tubeaxial fans use belt drives to achieve fan speeds below 1,100
- well-suited for ducted HVAC installations
- ventilation applications
Vaneaxial Fans
(see Fig. E.9)
- a vaneaxial fan is essentially a tubeaxial fan with outlet vanes to straighten the airflow, converting the airstream’s kinetic energy to static pressure
- the airflow profile is uniform
- when equipped with variable pitch blades, can be adjusted to change the angle of attack to the incoming air stream and air delivery rate
- have unstable regions to the left of the peak pressure
- highly efficient: when equipped with airfoil blades and built with small clearances, efficiencies up to 85 percent are achievable
- usually directly attached to the motor shaft
- typically used in medium- to high- pressure applications, such as induced draft service for a boiler exhaust
- low rotating mass, which allows them to achieve operating speed relatively quickly
- emergency ventilation
- reversal of air flow direction
TYPES OF AXIAL FANS
Fig. E.6
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The performance curve for propeller fans is shown in Figure E.7.
PROPELLER FAN
Fig. E.7
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The performance curve for tube axial fans is shown in Figure E.8.
TUBEAXIAL FAN
Fig. E.8
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Vane axial fans, a refinement over the axial fan involving flow straightening vanes, have a characteristic curve as shown in Figure E.9.
VANEAXIAL FAN
Fig. E.9
All three types of axial fans exhibit the stalling characteristic (where the pressure drops as the flow increases to the left of the peak pressure region) so care must be exercised in their application.
c. Special Fan Designs Bifurcated Fans
• In industrial processes that require the extraction of sticky, corrosive or volatile fumes, specially designed direct drive axial fans can be used.
• The motor of the axial fan is equipped with a unique protective casing that allows the motor to be removed from the airstream while maintaining a direct-drive arrangement. The casing protection is normally made from plastic or coated metal.
• The mating flanges at each end of the casing are identical. The casing diameter however is increased in barrel fashion around the casing to permit smooth passage of a similar cross-section of air concentric with the motor enclosure.
Centrifugal Inline Fans Centrifugal fans are used in commercial applications where high efficiency, low sound levels and space are the main considerations. They have a direct-drive or belt-driven airfoil or backward inclined impellers, mounted perpendicular in a rectangular or tubular casing with ample clearance around the blade tips.
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The air is discharged radially from the blade tips and must turn 90 degrees to pass through the fan exit, which is in line with the impeller inlet. This fan produces what is called ‘‘mixed flow.’’ Centrifugal Roof Exhausters
• These specialized fans are designed as a package together with their housing. They are designed to exhaust air to the outdoors.
• The down-discharge configuration is used for exhausting relatively clean air. The up-blast configuration is used for exhausting hot or contaminated air.
• They are usually direct-driven or belt-driven airfoil or backward inclined impellers in a multi-component housing.
• The housing is comprised of a curb cap with an integral inlet venturi, a shroud with drive-mounting support and a weatherproof motor hood.
• The impeller has an inlet cone that allows mixed flow through the impeller blade passages and air exits radially from the blade tips through a concentric discharge passage. The shroud redirects the air, either down or up.
Utility Fans Packaged utility fans, complete with the motor (direct or belt driven), are available for commercial and industrial ventilation applications requiring low to medium air volumes and pressures. These fans are usually equipped with forward curved or backwards inclined blades.
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APPENDIX F
TYPICAL FAN & BLOWER DATA SHEET
ASSALUYEH AMMONIA/ UREA COMPLEX
OWNER: PETROCHEMICAL INDUSTRIES DEVELOPMENT MANAGEMENT COMPANY
Dec. No.: 10-DSH-RE-005
DATA SHEET FOR CEVTRIFUGAL BLOWER Rev. 0
1. GENERAL 1.1 OWNER: PIDMCO SITE: Bandar Assaluyah, lran
1.2 UNIT: 2050 MTPD Ammonia Unit ITEM NO.: C-1003(101-BJ2) 1.3 TYPE OF FAN: Centrifugal(Backward) SERVICE: Induced Draft Fan
1.4 DU1Y: Continuous AREA CLASSIFICATION: Unclassified 1.5 TYPE OF DRIVER: Variable Speed Turbine
1.6 CONTINUOUS OPERATION: hour YES WITHOUT OVERHAUL: hour 1.7 INSTALLATION: INDOOR OUTDOOR MODE OF OPERATION:
2. OPERATING CONDITIONS 2.1 GAS COMPOSITION: Mole % composition: 66.87% Nitrogen, 1.57% Oxygen,7.39 C02, 24.17% Water 2.2 CAPACITY AT SUCTION m3/hr (Actual) NORMAL: RATED DESIGN: MINIMUM:
2.3 CAPACITY N m3/hr WET: NORMAL: Nate 10 RATED DESIGN: Note10 MINIMUM: Note 10
2.4 SUCTION PRESSURE mbarg DESIGN: Note 10 NORMAL :Note 10 MINIMUM: Note 10
2.5 DISCHARGE PRESSURE mbarg DESIGN: Note 10 NORMAL Note 10 MINIMUM: Note 10
2.6 TOTAL OUTLET PRESS. DIFF. mbar DESIGN: Note 10 NORMAL :Note 10 MINIMUM: Note 10
2.7 PRESS.LOSSE ACCESSORIES mbar STATIC DIFF.PRESS mbar.: By vendor TOTAL PRESS mbar.: By vendor
2.8 SUCTION TEMP. °C: DESIGN: Note 10 NORMAL: Note 10 MINIMUM: Note 10
2.9 DISCHARGE TEMP. °C: DESIGN: By vendor NORMAL: By vendor
3. PERFORMANCE DATA 3.1 CAPACITY AT SUCTION m3/hr: DISCHARGE-PR.- STATIC -mbar:
3.2 CAPACITY N m3/hr: (2) DIFFERENTIAL PR.: mbar: 3.3 STATIC EFFICIENCY: % SPEED: 1500 (Note11) RPM 3.4 BKW NORMALK.WTS: 1250@Maximum (Note 11) NOISE LEVEL: SEE NOTE 7 3.5 @MIN. SUC. TEMP. K.WTS: PERFORMANCE CURVE NO.:
3.6 ALLOWABLE WORKING TEMP. °C: ALLOWABLE WORKING PRESS: CORROSION ALLOWANCE: 3mm on CS
4. CONSTRUCTION DATA 4.1 MANUFACTURER: MODEL: 4.2 SUCTION: SINGLE DOUBLE CASING SPLIT: FLANGE :RATING I SIZE
4.3 TYPE OFIMPELLER : SIZE OF IMPELLER ,mm DIA.
4.4 JOURNAL BEARING :TYPE I SIZE THRUST BEARING TYPE/SIZE LUBRICATION: OIL GREASE
4.5 BEARING COOLING REQD. YES NO , if required by vendor. 4.6 SHAFT SEAL: IPACKING LABYRINTH MAKE: SIZE : CODE: 4.7 TYPE OF DRIVE: DIRECT GEAR V-BELT 4.8 GD2 AT DRIVE SHAFT END: ROTATION VIEWED FROM COUPLING END CW CCW
4.9 DAMPER TYPE: INLET OUTLET DAMPER CONTROL: AUTOMATIC DMANUAL
4.10 INLET GUIDE VANES: YES NO IGV CONTROL: AUTOMATIC MANUAL TYPE OF ACTUATORS: CONTROL SIGNAL:
4.11 SUCTION FILTER YES NO TYPE: DEGREE OF FILTRATION 4.12 VIBRATIONELIMINATORS YES NO FOR: NUMBER: MAKE: 4.13 COMMON BASE PLATE YES NO NOZZLE ORIENTATION: (SEE SHEET 2 OF 4) 4.14 COUPLING: 5. MATERIALS (DESIGNATION AS PER ASTM) 5.1 CASING: SEE NOTE 6 DAMPER / IGV:
5.2 IMPELLER HUB DISK: SEE NOTE 6 IMPELLER BLADE: SEE NOTE 6
5.3 IMPELLER COVER DISK: SHAFT: SEE NOTE 6 5.4 BASEPLATE: SHAFT SLEEVE: 5.5 SUCTION FILTER ELEMENT I HOUSING: BEARING PEDESTAL: 5.6 MOTOR PEDESTAL: BASEFRAME: 5.7 COUPLING HUB: COUPLING SPACER:
(to be continued)
July 2014
IPS-E-PR-755(1)
40
APPENDIX F (continued)
ASSALUYEH AMMONIA/ UREA COMPLEX
OWNER: PETROCHEMICAL INDUSTRIES DEVELOPMENT MANAGEMENT COMPANY
Dec. No.: 10-DSH-RE-005
DATA SHEET FOR CEVTRIFUGAL BLOWER Rev. 0
6. INSPECTION AND TESTS
MATERIAL TEST CERTIFICATES FOR FOLLOWING COMPONENTS 6.1 IMPELLER SHAFT CASING SHAFT SLEEVE
6.2 RADIOGRAPH TEST FOR SEE APPENDIX-C of 00-ESS-RE-305
6.3 ULTRASONIC & MAGNA FLUX TEST FOR SEE APPENDIX-C of 00-ESS-RE-305
6.4
6.5
INDICATES ONLY CERTIFICATE AND REQUIRED INDICATES TEST TO BE WITNESSED
BY PURCHASERS INSPECTOR
7. INSPECTION TESTING
7.1 ITEM REQ’D A B
7.2 MILL TEST REPORT
7.3 HYDROSTATIC TEST
7.4 NON DESTRUCTIVE EXAMINATION
7.5 BALANCING TEST OF ROTOR
7.6 CLEARANCE CHECK
7.7 PERFORMANCE TEST
7.8 MECHANICAL RUNNING TEST
7.9 SOUND LEVEL TEST
7.10 CHECK BEARING AFTER MECHANICAL RUNNING TEST
7.11 VISUAL AND DIMENSIONAL INSPECTION
7.12 MOTOR TEST
7.13 HYDROSTATIC TEST OF LUBE OIL UNIT
7.14 SHOP RUNNING TEST FOR LUBE OIL UNIT
7.15 VISUAL AND DIMENSIONAL INSPECTION FOR LUBE OIL UNIT
NOTE: A=TO BE WITNESSED B=NOT TO BE WITNESSED
NOZZLE ORIENTATION VIEWING FROM COUPLING END
(to be continued)
July 2014
IPS-E-PR-755(1)
41
APPENDIX F (continued)
ASSALUYEH AMMONIA/ UREA COMPLEX
OWNER: PETROCHEMICAL INDUSTRIES DEVELOPMENT MANAGEMENT COMPANY
Dec. No.: 10-DSH-RE-005
DATA SHEET FOR CEVTRIFUGAL BLOWER Rev. 0
8. GAS COMPOSITION AND PROPERTY
8.1 SOLID COMPONENTS YES NO
8.2 TOXIC YES NO
8.3 FLAMMABLE YES NO
8.4 CORROSIVE YES NO
8.5 EROSIVE YES NO
8.6 ADHESIVE YES NO
8.7 COMPONENTS CHEMICAL SYMBOL
MOLAR MASS kg/kmol Refer Details Given in Operating Conditions on Sheet 1of 4.
8.8 CARBON DIOXIDE
8.9 OXYGEN
8.10 NITROGEN
8.11 WATER
8.12
8.13 MOLAR MASS,TOTAL kg/kmol 26.8
8.14 SPECIFIC HEAT CAPACITY kj/(kg°C)
8.15 VISCOSITY cp 0.02
8.16 SPECIFIC WEIGHT kg/m3
8.17 ISENTROPIC EXPONENT
8.18 COMPRESSIBILITY
8.19 WATER CONTENT kg/kg
8.20 RELATIVE HUMIDITY 76% in summer, 74% in winter
(to be continued)
July 2014
IPS-E-PR-755(1)
42
APPENDIX F (continued)
ASSALUYEH AMMONIA/ UREA COMPLEX
OWNER: PETROCHEMICAL INDUSTRIES DEVELOPMENT MANAGEMENT COMPANY
Dec. No.: 10-DSH-RE-005
DATA SHEET FOR CEVTRIFUGAL BLOWER Rev. 0
REMARKS: 1) Fan to be mechanically designed for maximum flue gas temperature of 390 Deg.C.
2) Fan to be supplied with acoustic insulation, if required.
3) Site barometer: 990 - 1100 mbar
4) The critical speed shall be at least 35% above the maximum continuous speed.
5) Utilities Available-
Plant Utility Air- Operating Pressure/Temperature: 7.5 barg/Ambient.
Design Pressure/Temperature: 10.5 barg/60 Deg.C.
Cooling Water- Operating Pressure/Temperature: 4.5 barg/38 Deg.C
Design Pressure/Temperature: 7.5 barg/100 Deg.C(max).
Maximum allowable cooling water temperature rise -10 Deg.C
6) Anti-Dew point corrosion material to be considered. Expected maximum dew point is 103 Deg.C
7) The noise level of less than 85 dB (A) at 1m of the equipment.
8) Suction and discharge expansion joints shall be provided by vendor.
9) No copper and copper alloys are permitted.
10) OPERATING CONDITIONS
Case Unit Minimum Normal Maximum Start-Up Capacity Nm3/hr 152100 304335 380450 151000
Molecular Weight 26.8 26.8 26.8 28.11 Pressure Suction
(Static) mbarg -8.8 -35.3 -56.9 -8.8
Pressure Discharge (Static) mbarg 0 0 0 0
Total Outlet Differential Pressure mbarg 8.8 35.3 56.9 8.8
Temperature Deg.C 144-390 144 158 Ambient (5 to 48) Gas Analysis See Sheet 1 See Sheet 1 See Sheet 1 Air
11) Vendor to verify
