TCE DSIGN GUIDE.pdf

TATA CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-319

DESIGN GUIDE FOR HVAC SYSTEM

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FOR GREEN BUILDINGS

DESIGN GUIDE FOR HVAC SYSTEM FOR GREEN BUILDINGS

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DESIGN GUIDE FOR HVAC SYSTEM

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FOR GREEN BUILDINGS

CONTENTS SL. NO.

SH. NO.

TITLE

1.0

INTRODUCTION

1

2.0

SCOPE

1

3.0

DESIGN REQUIREMENTS

1

4.0

REFERENCES

16

APPENDICES APPENDIX NO. 1 2A 2B 3

SH. NO.

TITLE REFRIGERANT IMPACT CALCULATION

17

BUILDING ENVELOPE REQUIREMENTS AS PER ECBC-2007

19

BUILDING ENVELOPE REQUIREMENTS AS PER ASHRAE 90.1.2007 ASSEMBLY U FACTORS & SHGC FOR UNLABELED VERTICAL FENESTRATION & SKYLIGHTS

20 21

4

EQUIPMENT EFFICIENCIES AS PER ECBC-2007

22

5

EQUIPMENT EFFICIENCIES AS PER ASHRAE 90.1.2007

23

5A

AIR CONDITIONERS & CONDENSING UNITS

23

5B

WATER CHILLING PACKAGES

25

5C

PACKAGE UNITS

26

5D

HEAT REJECTION EQUIPMENTS

28

5E

NON-STANDARD CENTRIFUGAL CHILLERS UPTO 150 TR

29

5F

NON-STANDARD CENTRIFUGAL CHILLERS FROM 151 TR TO 300 TR

30

5G

NON-STANDARD CENTRIFUGAL CHILLERS BEYOND 301 TR

31

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DESIGN GUIDE FOR HVAC SYSTEM

SHEET iii OF iv

FOR GREEN BUILDINGS

6

PSYCHROMETRIC CHART (FREE COOLING SYSTEM)

32

7

FAN POWER LIMITATION PRESSURE DROP ADJUSTMENT

33

8

DUCT INSULATION

34

9

PIPING INSULATION

35

10

MINIMUM DUCT SEAL LEVEL

36

11

OUTDOOR AIR FLOW CALCULATION

37

12 A

MINIMUM VENTILATION RATES

38

12 B

MINIMUM EXHAUST RATES

40

12 C

ZONE AIR DISTRIBUTION EFFECTIVENESS

41

12 D

SYSTEM VENTILATION EFFICIENCY

42

13

MINIMUM SEPARATION DISTANCE

43

14

LIGHTING POWER DENSITIES

44

15

VOC LIMITS

47

16

VENTILATION RATE PROCEDURE (VRP) CALCULATION

48

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DESIGN GUIDE FOR HVAC SYSTEM

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REVISION STATUS REVISION NO.

DATE

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First issue

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FOR GREEN BUILDINGS 1.0

INTRODUCTION The built environment has a profound impact on our natural environment, economy, health, and productivity. Breakthroughs in building science, technology, and operations are now available to designers, builders, operators, and owners who want to build green and maximize both economic and environmental performance. The green building movement offers an unprecedented opportunity to respond to the most important challenges of our time, including global climate change, dependence on non sustainable and expensive sources of energy, and threats to human health. Green building is the practice of increasing the efficiency with which buildings use resources -- energy, water, and materials — while reducing building impacts on human health and the environment during the building's lifecycle.

2.0

SCOPE This document gives guidelines for design of HVAC system for Green Buildings.

3.0

DESIGN REQUIREMENTS Following requirements should be met for the HVAC system for Green Buildings to acquire various prerequisites & credits as per LEED (Leadership in Energy and Environmental Design) rating system:

3.1

REFRIGERANTS

3.1.1

Minimum requirement: Zero use of CFC based refrigerants in new building HVAC & R systems. In case of renovations or using existing HVAC & R equipments, identify existing equipments that uses CFC refrigerants & schedule for replacement of these refrigerants.

3.1.2

As per Indian Green Building Council (IGBC)-LEED rating system, equipment that do not contain HCFCs should be used for getting additional credits. Whereas as per LEED-USGBC HVAC&R equipment must comply with the following formula, which sets a maximum threshold for the combined contributions to ozone depletion and global warming potential: LCGWP + LCODP x 105

100

Where LCODP LCGWP LCODP LCGWP GWPr ODPr Lr

= [ODPr x (Lr x Life +Mr) x Rc] / Life = [GWPr x (Lr x Life + Mr) x Rc] / Life : Lifecycle Ozone Depletion Potential (lb CFC 11/Ton-Year) : Lifecycle Direct Global Warming Potential (lb CO2/Ton-Year) : Global Warming Potential of Refrigerant (0 to 12,000 lb CO2/lbr) : Ozone Depletion Potential of Refrigerant (0 to 0.2 lb CFC 11/lbr) : Refrigerant Leakage Rate (0.5% to 2.0%; default of 2% unless otherwise demonstrated) ISSUE R0

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FOR GREEN BUILDINGS Mr

: End-of-life Refrigerant Loss (2% to 10%; default of 10% unless otherwise demonstrated) : Refrigerant Charge (0.5 to 5.0 lbs of refrigerant per ton of gross ARI rated cooling capacity) : Equipment Life (10 years; default based on equipment type, unless otherwise demonstrated)

Rc Life

In case of multiple type of equipment average of all equipment must be calculated using following formula : (LCGWP + LCODP x 105 ) x Qunit

100

QTotal Where Qunit

: ARI rated cooling capacity of individual refrigeration / HVAC unit (TR)

QTotal

: Total ARI rated cooling capacity of all refrigeration / HVAC unit (TR)

Sample Calculation for the same is shown in APPENDIX 1. 3.1.3

Select HVAC&R equipment with reduced refrigerant charge and increased equipment life. Maintain equipment to prevent leakage of refrigerant to the atmosphere.

3.2

BUILDING ENVELOPE

3.2.1

The exterior building envelope shall be designed with either Residential or Non-residential requirements in APPENDIX 2, Building Envelope Requirements, (excerpts from American Society of Heating, Refrigerating and Air-Conditioning Engineers (ASHRAE) 90.1.2007 & Energy Conservation Building Code (ECBC)). As per green building requirements stringent requirements of the above (ASHRAE or ECBC) shall be followed. For all opaque surfaces, compliance to standard shall be demonstrated by either Minimum rated R values of insulation or Maximum U factor.

3.2.2

The total vertical fenestration area shall be less than 40% (as per ASHRAE 90.1.2007) of the gross wall area. In ECBC vertical fenestration area limit is up to 60% of the gross wall area. The total skylight area shall be less than 5% of the gross roof area as per ASHRAE 90.1.2007 and ECBC.

3.2.3

Design building envelope so that space conditions are maintained within the range specified in ASHRAE standard 55-2004, Thermal comfort Conditions for Human Occupancy. Documentation shall be as under (Excerpts from ASHRAE standard 55-2004, Section 6.1.1) (i) The design criteria of the system in terms of indoor temperature, humidity, including any tolerance based on stated design outdoor ambient conditions and total indoor loads should be stated. ISSUE R0

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FOR GREEN BUILDINGS (ii) The system input or output capacities necessary to attain the design indoor conditions at design outdoor ambient conditions should be stated along with system supplied and installed capacities. (iii) The limitations of the system to control the environment of the zone(s) should be stated. (iv) The overall space supplied by the system should be shown in a plan view layout, with all individual zones within it identified. All terminal units should be shown with type and flow. (v) Items affecting indoor comfort (such as decor, significant structural items) should be shown. Notes should be provided to identify areas and location within space relative to grilles, diffuser, sensors which should not be obstructed. (vi) Areas within any zone that lie outside the comfort area, where people should not be permanently located, should be identified. (vii) Location of all occupant adjustable controls should be identified and each should be provided with a legend describing which zone(s) and function it controls, how it is to be adjusted, the range of effect it can have and recommended setting for various times of day, season or occupancy load. (viii) A block diagram control schematic should be provided with sensors, controls and actuators identified for each zone. (ix) The general maintenance, operation and performance of the building systems should be state, followed by more specific comments on maintenance and operation of automatic controls and manually adjustable controls and the response of the system to each. Where necessary, specific seasonal settings of manual controls should be stated and major system changeovers that are required to be performed by a professional service agency should be identified. (x) Specific limits in the adjustment of manual controls should be stated. Recommendations for seasonal settings on these controls should be stated. A maintenance and inspection schedule for all thermal environmental related systems should be provided. (xi) Assumed electrical load for lighting and equipment in occupied spaces (including diversity considerations) used in HVAC load calculations should be documented along with any other significant thermal and moisture loads assumed in HVAC load calculations and any other assumptions based upon which HVAC and control design is based. 3.2.4

Fenestration and Doors Where fenestration and doors are used in the building envelope, it shall comply with following requirements:

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FOR GREEN BUILDINGS 3.2.4.1 The U factor, SHGC & air leakage rate shall be determined by a laboratory accredited by a nationally recognized accreditation organization such as National Fenestration Rating Council (NFRC). 3.2.4.2 U factors & SHGC for the overall fenestration product shall be determined in accordance with ISO 15099 as per ECBC or NFRC 100 & NFRC 200 respectively as per ASHRAE 90.1.2007. Or alternatively U factors & SHGC from APPENDIX 3, Assembly U factors & SHGC for unlabeled vertical fenestration & skylights, are acceptable. 3.2.4.3 U factors for unlabeled opaque doors shall be as under (a) Un-insulated single layer metal swinging or non-swinging doors, including single layer un-insulated access hatches : 1.45 Btu/hr ft2 °f (b) Un-insulated double layer metal swinging or non-swinging doors, including double layer un-insulated access hatches : 0.75 Btu/hr ft2 °f (c) Insulated metal swinging doors including fire rated doors, insulated access hatches : 0.50 Btu/hr ft2 °f (d) Wood doors , minimum nominal thickness of 1.75” including panel doors with minimum panel thickness of 1.125” , solid core flush doors, & hollow core flush doors :0.50 Btu/hr ft2 °f (e) Any other wood door : 0.60 Btu/hr ft2 °f 3.2.4.4 The U factor & the air leakage rate shall be identified on a permanent nameplate installed on the product by the manufacturer. 3.2.5

Air Leakage Rates

3.2.5.1 All openings in the building envelope such as joints around doors & fenestrations, corners, utility services penetrations, openings shall be sealed, gasketed or weather stripped to minimise air leakage. 3.2.5.2 Air leakage shall not exceed 1 cfm/ft2 for glazed swinging entrance doors & 0.4 cfm/ft2 for all other products. 3.2.6

Insulation Where insulation is used in the building envelope it shall comply with following requirements:

3.2.6.1 The rated R value of the insulation shall be clearly identified by the manufacturer on each piece of building envelope insulation 3.2.6.2 Insulation material shall be installed as per manufacturer’s recommendation, in substantial contact with inside surface & in such a manner so as to achieve the rated R value of the insulation.

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FOR GREEN BUILDINGS 3.2.6.3 Roof insulation shall not be installed on a suspended ceiling with removable ceiling panels. 3.2.6.4 Loose fill insulation shall not be used in attic roof spaces when slope of the ceiling is more than 3:12 3.2.6.5 Insulation material in ground contact shall have water absorption rate no greater than 3%when tested in accordance with ASTM C272 3.2.6.6 Flexible bat insulation installed in floor cavities shall be supported in a permanent manner by supports no greater than 24” on centre. 3.2.6.7 Any equipment, fixtures shall not be recessed in such a manner as to affect the insulation thickness. 3.2.6.8 Exterior insulation shall be covered with protective material to prevent damage from moisture, sunlight, wind, maintenance. 3.2.6.9 Insulation shall extend over the full component area to the required rated R value of insulation. 3.3

HVAC EQUIPMENTS

3.3.1

Equipment Efficiencies

3.3.1.1 HVAC equipments shall have minimum performance as per APPENDIX 4, Minimum Equipment Efficiency requirements as per ECBC-2007 or APPENDIX 5, Minimum Efficiency requirements as per Ashrae 90.1.2007, whichever is stringent, at specified rating conditions when tested in accordance with specified test procedure. Where multiple rating conditions or performance requirements are provided, the equipment shall satisfy all rated conditions. APPENDIX details are as under : APPENDIX 4

Minimum Efficiency Requirement As per ECBC

APPENDIX 5

Minimum Efficiency Requirement as per ASHRAE 90.1.2007 as under

APPENDIX 5 A

Air Conditioners & Condensing Units

APPENDIX 5 B

Water Chilling Packages

APPENDIX 5 C

Package Units

APPENDIX 5 D

Heat Rejection Equipments

Water cooled Centrifugal Chillers that are not designed for operation at ARI standard 550/590 test conditions shall have performance requirements (full load COP & NPLV) as per following APPENDICES: APPENDIX 5 E

Chiller Capacity < 150 TR ISSUE R0

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FOR GREEN BUILDINGS APPENDIX 5 F

150 TR

Chiller Capacity < 300 TR

APPENDIX 5 G

Chiller Capacity

300 TR

The table values are applicable over the following full load design range : Leaving Chilled water temperature

:40°F to 48°F

Entering Condenser water temperature

:75°F to 85°F

Condenser water temperature rise

:5°F to 15°F

3.3.1.2 Equipment efficiencies information provided by the manufacturer shall be verified by certified programme & data furnished by manufacturer 3.3.1.3 Equipments shall carry a permanent label installed by manufacturer stating that the equipment complies with the requirements of ASHRAE standard 90.1 3.3.2

Air and Water Economizers

3.3.2.1 Each individual cooling fan system that has a design supply capacity over 2500 cfm and a total mechanical cooling capacity over 6.3 TR shall include either air side or water side economizer(as per para 5.3.1 of ECBC). Projects in Hot dry and Warm humid climate zones are exempt and individual ceiling mounted fan systems ( hmc (i)

The outdoor air damper is opened to a minimum position, taking minimum outdoor air required for ventilation.

(ii)

Return air damper is fully open and exhaust air damper is opened to a minimum position.

(iii)

Refrigeration plant is working to cool the mixing air to the supply air temperature.

Case II - When the enthalpy of outside air is equal to or lower than the mixing air enthalpy but more than the supply air enthalpy (hsa), hsa < hos ≤ hmc

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FOR GREEN BUILDINGS (i)

The outdoor air damper and exhaust air damper both are fully open.

(ii)

Return air damper is closed.

(iii)

Refrigeration plant is working to cool the outdoor air to the supply air temperature.

Case III - When the enthalpy of outside air is lower than the supply air enthalpy, hos ≤ hsa (i)

The outdoor air damper and Return air damper both are modulated to achieve the supply air temperature.

(ii)

Exhaust air damper is also modulated.

(iii)

Refrigeration plant is switched off

Careful analysis is required for designing & implementing the above scheme considering outside conditions throughout the year. 3.3.3

Each HVAC system having a total fan system motor nameplate hp exceeding 5 hp shall meet the following :

3.3.3.1 Each HVAC system at fan system design conditions shall not exceed the allowable fan system motor nameplate hp (Option 1) or fan system bhp (Option 2) as shown below Limit

Constant Volume

Variable Volume

Option1

Allowable nameplate Motor HP

hp ≤ CFM*0.0011

hp ≤ CFM*0.0015

Option2

Allowable fan system BHP

bhp ≤ CFM*0.00094+A

bhp ≤ CFM*0.0013+A

where CFM is the maximum design supply air flow rate to conditioned spaces served by the system in cfm and A is the sum of (PD x CFM (through each device) / 4131). PD shall be as given in the APPENDIX 7 3.3.3.2 The selected fan motor shall be no larger than the first available motor size greater than the bhp. 3.3.3.3 Individual VAV fans with motors 10 hp or larger shall meet one of the following requirement (a)

The fan shall be driven by a mechanical or electrical variable speed drive

(b)

The fan shall be a vane axial fan with variable pitch blades. ISSUE R0

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FOR GREEN BUILDINGS (c)

3.3.4

The fan shall have other controls and devices that will result in fan motor demand of no more than 30% of design wattage at 50% of design air volume when static pressure set point equals one third of the total design static pressure based on manufacturers certified fan data.

HVAC hydronic systems having a total pump power exceeding 10 hp shall meet following provisions

3.3.4.1 HVAC pumping systems that include control valves designed to modulate or step open and close as a function of load shall be designed for variable fluid flow and shall be capable of reducing pump flow rates to 50% or less of the design flow rate. Individual pumps serving variable flow systems having a pump head in excess of 100 ft and motor exceeding 50 hp shall have controls and/or devices (such as variable speed control) that will result in pump motor demand of no more than 30% of design wattage at 50% of design water flow. The devices shall be controlled as a function of desired flow or to maintain a minimum required differential pressure. Differential pressure shall be measured at or near the most remote heat exchanger requiring the greatest differential pressure. 3.3.4.2 When a chilled water plant includes more than one chiller, provisions shall be made so that the flow in the respective chiller can be automatically stopped when a chiller is shut down. 3.3.5

Each fan of heat rejection equipment (whose energy usage is not included in equipment efficiency tables listed in Sr. No. 3.3.1.1) of motor rating 7.5 hp or larger shall have the capability to operate at 2/3rd of full speed or less and shall have speed controls to control the leaving fluid or condensing temperature or pressure of the heat rejection equipment.

3.3.6

Individual fan systems that have both a design supply air capacity of 5000 cfm or greater and have a minimum outdoor air supply of 70% or greater of the design supply air quantity shall have an energy recovery system with at least 50% recovery effectiveness (i.e. Change in enthalpy of the outdoor air supply equal to 50% of the difference between the outdoor air and return air at design conditions).

3.3.7

Kitchen exhaust hoods larger than 5000 cfm shall be provided with makeup air sized for at least 50% of exhaust air volume.

3.3.8

Building with fume hood systems having a total exhaust rate greater than 15000 cfm shall include at least one of the following features

3.3.8.1 VAV hood exhaust and room supply systems capable of reducing exhaust and makeup air volume to 50% or less of design values 3.3.8.2 Direct make up air supply equal to at least 75% of the exhaust rate, cooler to no cooler than 3°F above room set point and heated no warmer than 2°F below room set point, no humidification added, no simultaneous heating and cooling used for dehumidification control.

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FOR GREEN BUILDINGS 3.3.8.3 Heat recovery system to precondition makeup air from fume hood exhaust in accordance with Sr. No. 3.3.6 3.3.9

Controls

3.3.9.1 Design the HVAC system (heating & cooling both) to provide individual comfort controls to allow adjustments to suit individual needs. Also provide comfort controls for all shared multi-occupant spaces to enable adjustments that meet group needs. Conditions of thermal comfort shall be as described in ASHRAE standard 55-2004. 3.3.9.2 Controls for metering the energy used shall be included in the design. 3.3.9.3 Temperature range or dead band for controls shall be 5°F except for special applications 3.3.9.4 When HVAC systems are not intended to operate continuously , HVAC system shall be designed with controls (i) that shall start and stop the system under different time schedule , (ii) an occupant sensor that is capable of shutting the system off when no occupant is sensed for a period of 30 minutes, (iii) a manually operated timer capable of bring adjusted to operate the system for upto two hours, (iv) an interlock with the security system that shuts the system off when the security system is activated. 3.3.9.5 Individual HVAC systems having capacity in excess of 10000 cfm shall have the optimum start control, the control algorithm as a minimum shall be a function of difference between space temperature and occupant set point and amount of time prior to scheduled occupancy. 3.3.9.6 HVAC systems serving zones that may operate non simultaneously shall be divided into isolated areas. Zones shall be grouped so that no individual zone exceeds 25000 ft2 of conditioned floor area. Controls shall be provided for automatic shutting off the system in particular zone when system is not in use. 3.3.9.7 Stair and shaft vents shall be equipped with motorized dampers that are capable of being automatically closed during normal building operation & are interlocked to open when required by fire and smoke detection system. All outdoor supply air and exhaust air systems shall be equipped with motorised dampers that will automatically shut down when the spaces served are not in use. 3.3.9.8 Outdoor air supply & exhaust dampers shall have a maximum leakage rate not more than 4 cfm / ft2 @ 1” SP for motorized dampers & no leakage at 1” SP for non motorized damper. However damper smaller than 24” in either dimension may have leakage of 40 cfm/ft2 3.3.10 Insulation 3.3.10.1 All supply & return ductwork insulation shall be in accordance with APPENDIX 8, Duct Insulation ISSUE R0 FILE NAME: F120R3.DOC

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FOR GREEN BUILDINGS 3.3.10.2 All piping insulation shall be in accordance with APPENDIX 9, Piping Insulation 3.3.10.3 Insulation exposed to weather shall be protected by aluminium, sheet metal, painted canvas or plastic cover. 3.3.11 Ductwork and plenum shall be sealed in accordance with APPENDIX 10, Minimum Duct Seal Level 3.3.12 Ductwork that is designed to operate at static pressures in excess of 3” w.c. shall be leak tested for 25% of the installed duct area. The maximum permitted leakage shall be Lmax = CL x P0.65 where Lmax

=Maximum permissible leakage , cfm / 100 ft2 duct surface area

CL

=Duct leakage class, cfm / 100 ft2 at 1” W.C. 6 for rectangular sheet metal / fibrous & round flexible ducts 3 for round/flat oval sheet metal / fibrous glass duct

P

=test pressure (design duct pressure)

3.3.13 Completion Requirements Following construction documents shall be provided to building owner or representative within 90 days of acceptance of the system: (a)

As built drawings

(b)

Operation & Maintenance Manual

(c)

Air & Water balancing reports

3.3.14 HVAC Electrical equipment & cabling (a) All Motors shall have a minimum acceptable nominal full load motor efficiency not less than IS 12615 standard for energy efficient motors (b) Feeder conductors shall be sized for a maximum voltage drop of 2% at design load. Branch circuit conductors shall be sized for a maximum voltage drop of 3% at design load 3.4

VENTILATION FOR ACCEPTABLE INDOOR AIR QUALITY

3.4.1

Outdoor Air Quality

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FOR GREEN BUILDINGS 3.4.1.1 Outdoor air quality shall be evaluated for the project site prior to completion of ventilation system design for checking the outdoor air contamination levels. A survey of the building site & its immediate surroundings shall be conducted during hours the building is expected to be normally occupied to identify local contaminants from surrounding facilities. The observations of the survey shall be documented (as per ASHRAE 62.1.2004) & discussed with the building owner or their representatives. 3.4.1.2 If the outdoor air is judged to be unacceptable in accordance with ASHRAE 62.1.2004, ventilation system that provides outdoor air through a supply fan shall comply with the following : (a) Air filters having MERV of 6 or higher shall be provided (b) Air cleaning devices having minimum volumetric ozone removal efficiency of 40% for ozone shall be provided when outdoor ozone levels are expected to exceed 0.16 ppm. Air cleaning devices for ozone are not required when outdoor air intake results in 1.5 ACPH or less 3.4.2

Mechanical Ventilation

3.4.2.1 As a minimum requirement for green buildings, mechanical ventilation systems shall be designed using the Ventilation Rate Procedure (VRP) or the applicable local code whichever is more stringent. Refer APPENDIX 11, for Ventilation Rate Procedure details. Further additional credit points can be acquired by increasing the ventilation rates to all occupied spaces by at least 30% above minimum rates required by ASHRAE standard 62.1.2004. Minimum ventilation rates are given in APPENDIX 12 A. One example is given in APPENDIX 16. 3.4.2.2 Various tables required for calculation in ventilation rate procedure are given as under : APPENDIX 12 A

-

Minimum Ventilation Rates for Breathing Zone

APPENDIX 12 B

-

Minimum Exhaust Rates

APPENDIX 12 C

-

Zone Air Distribution Effectiveness

APPENDIX 12 D

-

System Ventilation Efficiency

3.4.2.3 Demand control ventilation (DCV) is required for spaces larger than 500 ft2 and with design occupancy for ventilation of greater than 40 people per 1000 ft2 of floor area and served by systems with either an air side economizer or automatic modulating control of the outside air damper or both.

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FOR GREEN BUILDINGS DCV is a system that adjusts the amount of outside air based on the number of occupants and the ventilation demands that those occupants create. DCV systems uses CO2-based sensors, which measure the buildup of CO2 from the occupants present and control the outside air flow. In this system CO2 sensors monitor CO2 levels in the air inside a building or occupied areas, and an air-handling system uses this data from the sensors to regulate the amount of ventilation air admitted. 3.4.2.4 Use heat recovery, where appropriate, to minimize the additional energy consumption associated with higher ventilation rates. 3.4.2.5 Indoor Air Quality (IAQ) procedure can be used as an alternative to VRP (described at Sr. No. 3.4.2.1) as per Ashrae standard 62.1. IAQ procedure allows credit to be taken for controls that remove contaminant. It is also used where the design is intended to attain specific target contaminant concentrations or levels of acceptability of perceived indoor air quality. Air purification equipment are used in the HVAC systems enabling a reduction in ventilation air flow rates from the levels required by VRP. 3.4.2.6 Use of controlled injection of ozone can help reduce the quantity of fresh air. Ozone is a powerful oxidant, which removes odour, Volatile Organic Compounds (VOC) and even fungi by oxidation. This reduces the oxygen requirement in the form of ventilation and air is mainly required for diluting CO2. With ozone generators, ozone is injected into the central air conditioning system (ducts) & concentration is controlled by sensors activated by excessive VOC concentrations & turned off by excess ozone concentration sensors. Ozone concentration is required to be kept below harmful limit (0.05 ppm as per Ashrae). 3.4.3

Natural Ventilation

3.4.3.1 Naturally ventilated spaces shall be permanently open to and within 8 m of operable wall or roof openings to the outdoors. Openable area shall be minimum of 4% of the net occupiable floor area. 3.4.3.2 An engineered natural ventilation system when approved by the authority having jurisdiction need not meet the above requirement 3.4.3.3 Credits may be acquired by designing natural ventilation system to meet recommendations set forth in the Chartered Institution of Building Services Engineers (CIBSE) Application manual 10:2005, follow flow diagram process shown in figure 2.8 of CIBSE Application Manual 10 OR 3.4.3.4 Follow 8 design steps, as under, described in Carbon Trust Good Practice Guide 237 (a) Develop design requirements (b) Plan air flow paths (c) Identify building uses & features that require special attention ISSUE R0 FILE NAME: F120R3.DOC

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FOR GREEN BUILDINGS (d) Determine ventilation requirement (e) Estimate external driving pressures (f) Select types of ventilation devices (g) Size ventilation devices (h) Analyze the design OR Use public domain software such as NIST’s CONTAM, Multizone modelling software, along with LoopDA, Natural ventilation sizing tool, to analytically predict room by room airflows. 3.4.4

Smoking rooms designed to contain, capture & remove environmental tobacco smoke (ETS) from the building must be directly exhausted to the outdoors, away from air intakes and building entry paths with no recirculation of ETS containing air. Exhaust shall be sufficient to create negative pressure differential with the surrounding spaces of at least an average of 0.02” of water gauge & minimum of 0.004” of water gauge when doors to smoking rooms are closed. Corridors common to ETS & ETS free areas shall be supplied with outdoor air at the rate of 0.1 cfm/ft2. Differential pressure generally used is 0.05” WC.

3.4.5

Sufficiently exhaust each space where hazardous gases or chemicals may be present or used to create negative pressure with respect to adjacent spaces when the doors to the room are closed. Exhaust rate must be at least 0.5 cfm / ft2 with no air recirculation. The pressure differential with the surrounding spaces must be at least an average of 0.02” & minimum of 0.004” when doors to the rooms are closed. Filters having MERV of 13 or higher must be provided for mechanically ventilated buildings. This requirement is used to acquire credit points and is not a mandatory requirement.

3.4.6

Minimum outdoor air intake shall be greater than the design maximum exhaust airflow when the mechanical air-conditioning are dehumidifying.

3.4.7

ASHRAE Standard 62.1 includes a design requirement intended to limit relative humidity in zones. To comply with this, designers must analyze the performance of the proposed air conditioning system at relatively severe latent load conditions, namely, with outdoor air at dehumidification design condition (design dew point and mean coincident dry-bulb temperature) and with no zone sensible-heat gain due to solar load. This analysis tests the dehumidification capability of the HVAC system configuration and control. It must show that zone relative humidity does not exceed 65% RH at these conditions. Some systems in some buildings in some climates can meet this requirement without direct humidity control, e.g., VAV systems that supply cool, dry primary air at all conditions. Other systems, however, such as traditional single-zone constant volume systems, supply warmer, moister air at part-sensible load and cannot maintain 65% RH or less without some enhancement. To comply with Standard 62.1, these systems must be reconfigured, to limit relative humidity indirectly, or using a zone humidistat and local reheat. ISSUE R0

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FOR GREEN BUILDINGS 3.4.8

Location of outdoor air intakes shall be such that the shortest distance from the intake to any specific potential outdoor contaminant source shall be equal to or greater than the separation distance listed in APPENDIX 13, Minimum Separation distance

3.4.9

Outdoor air intakes that are part of the mechanical ventilation system shall be designed to manage Rain entrainment, Rain intrusion, Snow entrainment in accordance with the ASHRAE standard 62.1.2004

3.4.10 Air shall be classified and its recirculation shall be limited in accordance with the following : (a) Class 1 air : Air with low contaminant concentration, low irritation intensity and inoffensive odour, may be recirculated or transferred to any space. (b) Class 2 : Air with moderate contaminant concentration, mild irritation intensity and mild offensive odour, may be recirculated within the space of origin or transferred or recirculated to other class 2 or class 3 or class 4 spaces utilized for similar purpose. (c) Class 3 : Air with significant contaminant concentration, irritation intensity and offensive odour, may be recirculated within the space of origin. Shall not be transferred or recirculated to any other space. (d) Class 4 : Air with highly objectionable fumes or gases & at harmful concentrations, Shall not be transferred or recirculated to any other space nor recirculated within space of origin. Classification of air leaving each space shall be in accordance with APPENDIX 12 A, Minimum Ventilation Rates for Breathing Zone 3.4.11 Mechanical ventilation systems shall include controls, manual or automatic, that enable fan system to operate whenever the spaces served are occupied. The system shall be designed to maintain the minimum outdoor airflow under any load condition. 3.4.12 The system may be designed to reset the design outdoor air intake flow as operating conditions change, these conditions include but are not limited to : 3.4.12.1Variation in occupancy or ventilation rate or CO2 levels (DCV as described at Sr. No. 3.4.2.3) in one or more individual zones. 3.4.12.2 A higher fraction of outdoor air in the supply due to intake of additional outdoor air for free cooling or exhaust air make up. 3.4.12.3 All air stream surfaces in equipments and ducts shall be designed and constructed in accordance with ASHRAE standard 62.1.2004, Resistance to mould growth & erosion 3.5

LIGHTING POWER DENSITIES

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FOR GREEN BUILDINGS Lighting power densities must not exceed the densities for the classified zone given in ANNEXURE 12, Lighting Power Densities 3.6

LOW EMITTING MATERIALS

3.6.1

Specifications shall clearly state the VOC limits as given in ANNEXURE 13, VOC Limits, for coatings, adhesives & paints that may be used in construction / installation of HVAC system. All adhesives and sealants must comply with South Coast Air Quality Management District Rule 1168

3.7

BUILDING PERFORMANCE As a minimum requirement for Green building, at least one of the following requirements shall be met.

3.7.1

LEED India requires establishing the minimum level of energy efficiency for the base building and systems and requires the building project to comply with mandatory provisions and prescriptive provisions of ASHRAE standard 90.1.2004 which are described above in earlier sections. Project should also comply with the final version of ECBC. Whereas as per LEED (USGBC) rating system it is required to demonstrate a 10% improvement in the proposed building performance rating for new buildings, or a 5% improvement in major renovations to existing buildings, compared with the baseline building performance rating. Baseline building performance rating is calculated according to the building performance rating method of ASHRAE Standard 90.1-2007 using a computer simulation model for the whole building project. Ashrae Standard 90.1-2007 requires that the energy analysis done for the building performance rating method include all energy costs associated with the building project. OR

3.7.2

Comply with the prescriptive measures of ASHRAE Advanced Energy Design Guide for Small Office Buildings 2004 / Small Retail Buildings 2006 / Small Warehouses & Self Storage Building 2008 or Core Performance Guide

3.7.3

Further, additional credits may be acquired by improving the performance of the proposed building (Ref 3.7.1). Credits are available up to 42% (48% as per LEED USGBC) improvement in performance of proposed building. This can be achieved by whole building energy simulation (using the Building Performance Rating Method in Appendix G of the ASHRAE standard 90.1) or prescriptive compliance path described in 3.7.2 above.

3.8

Green building design requirements also encourages use of onsite renewable energy such as Solar, Wind, Geothermal, Bio mass & Bio gas strategies. However this is not a mandatory requirement at present and is used to acquire additional credit points for Green building certification.

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FOR GREEN BUILDINGS

4.0

REFERENCES

4.1

LEED 2009 for New Construction and Major Renovations

4.2

Reference Guide for New Construction and Major Renovations (LEED-India NC) Version 1.0

4.3

ASHRAE Standard 90.1.2007, Energy Standard for Buildings except Low Rise Residential Buildings

4.4

ASHRAE Standard 62.1.2004, Ventilation for Acceptable Indoor Air Quality

4.5

Energy Conservation Building Code 2007

4.6

ASHRAE Standard 55-2004, Thermal comfort Conditions for Human Occupancy

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FOR GREEN BUILDINGS APPENDIX 1 REFRIGERANT IMPACT CALCULATION Sample calculation for arriving at the refrigerant impact is given here. Assumptions made for the calculation are as under Total no. of systems in complete project

:3 (System-1,2 and 3)

Refrigerant for Systems

:R-134a for System-1, 2 and 3

CALCULATION TABLE Description of system

System-1

System-2

System-3

R-134a

R-134a

R-134a

1000

1400

1400

GWPr

1320

1320

1320

ODPr

0

0

0

Refrigerant System Capacity

TR

Rc

lb

3300

3990

3990

Life

years

23

23

23

Lr

%

2

2

2

Mr

%

10

10

10

LCGWP

106059

128235

128235

LCODP

0

0.00

0.00

106059

128235

128235

LCGWP + LCODP x 105 Total impact

Total

3800

362529 95.4

Where LCODP LCGWP LCODP LCGWP GWPr ODPr

= [ODPr x (Lr x Life +Mr) x Rc] / Life = [GWPr x (Lr x Life + Mr) x Rc] / Life : Lifecycle Ozone Depletion Potential (lb CFC 11/Ton-Year) : Lifecycle Direct Global Warming Potential (lb CO2/Ton-Year) : Global Warming Potential of Refrigerant (0 to 12,000 lb CO2/lbr). Values taken from Refrigerant Reference Guide, National Refrigerants : Ozone Depletion Potential of Refrigerant (0 to 0.2 lb CFC 11/lbr). Values taken from Refrigerant Reference Guide, National Refrigerants ISSUE R0

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FOR GREEN BUILDINGS

Lr

: Refrigerant Leakage Rate (0.5% to 2.0%; default of 2% unless otherwise demonstrated)

Mr

: End-of-life Refrigerant Loss (2% to 10%; default of 10% unless otherwise demonstrated)

Rc

: Refrigerant Charge (0.5 to 5.0 lbs of refrigerant per ton of gross ARI rated cooling capacity). Values taken from equipment Manufacturer. : Equipment Life, Values taken from equipment Manufacturer

Life

Total impact

(LCGWP + LCODP x 105 ) x Qunit QTotal

100

Conclusion 1.0

Project satisfies minimum requirement of zero use of CFC refrigerant as specified in Sr. No. 3.1.1

2.0

Project will get the additional credit as it satisfies the requirement of total impact less than 100.

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FOR GREEN BUILDINGS APPENDIX 2 A BUILDING ENVELOPE REQUIREMENTS AS PER ECBC 2007 Walls & Roofs Climate Zones*

Roofs Composite Hot & Dry Warm & Humid Moderate Cold Walls Composite Hot & Dry Warm & Humid Moderate Cold Fenestrations Climate Zones*

24 Hour use buildings Hotels, Hospitals etc Assembly Insulation Max U Min R Value

Daytime use buildings Other Building Types Assembly Insulation Max Min R Value

Btu/hr ft2 °f

hr ft2 °f / Btu

Btu/hr ft2 °f

hr ft2 °f / Btu

0.0460 0.0460 0.0460 0.0715 0.0460

R-19.8 R-19.8 R-19.8 R-11.9 R-19.8

0.0715 0.0715 0.0715 0.0715 0.0715

R-11.9 R-11.9 R-11.9 R-11.9 R-11.9

0.0775 0.0775 0.0775 0.0759 0.0650

R-11.9 R-11.9 R-11.9 R-10.2 R-12.5

0.0775 0.0775 0.0775 0.0699 0.0620

R-11.9 R-11.9 R-11.9 R-11.4 R-13.3

Maximum U Factor

WWR

Maximum SHGC 40% 40%=WWR 60%

Btu/hr ft2 °f

Vertical Fenestration Composite Hot & Dry Warm & Humid Moderate Cold

0.58 0.58 0.58 1.22 0.58

0.25 0.25 0.25 0.40 0.51

0.20 0.20 0.20 0.30 0.51

Skylights Climate Zones* Composite Hot & Dry Warm & Humid Moderate Cold

Maximum U Factor With Curb W/o Curb Btu/hr ft2 °f

Btu/hr ft2 °f

1.98 1.98 1.98 1.98 1.98

1.36 1.36 1.36 1.36 1.36

Maximum SHGC 0 to 2% SRR 2.1 to 5% SRR

0.40 0.40 0.40 0.61 0.61

0.25 0.25 0.25 0.40 0.4

*Climate Zones as per ECBC , Climate Zone Map of India Definitions : WWR : Window to Wall Ratio, SRR : Skylight Roof Ratio

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FOR GREEN BUILDINGS APPENDIX 2 B BUILDING ENVELOPE REQUIREMENTS* AS PER ASHRAE 90.1.2007 Opaque Elements

Non-Residential Assembly Insulation Max U Min R Value

Residential Assembly Insulation Max Min R Value

Btu/hr ft2 °f

hr ft2 °f / Btu

Btu/hr ft2 °f

hr ft2 °f / Btu

Roofs Insulation above deck Metal Building Attic & Other

0.063 0.065 0.034

R-15 R-19 R-30

0.048 0.065 0.027

R-20 R-19 R-38

Walls, Above grade Mass Metal Building Steel framed Wood framed or other

0.58 0.113 0.124 0.089

NR R-13 R-13 R-13

0.151 0.113 0.124 0.089

R-5.7 R-13 R-13 R-13

Walls, Below grade

C-1.14

NR

C-1.14

NR

Floors Mass Steel Joist Wood framed & other

0.322 0.35 0.282

NR NR NR

0.322 0.35 0.282

NR NR NR

Fenestration 0 to 40% of wall Normal framing Metal framing (Curtainwall/Storefront)

1.2 1.2

1.2 1.2 SHGC 0.25 all

SHGC 0.25 all

*For Climate Zone 1 (A,B) & Prescriptive Building Envelope Option Notes : (1) Either above U values can be specified / used or alternatively U values for pre-calculated assemblies can be used from APPENDIX A of ASHRAE 90.1.2007. (2) For Min. R value, specifications listed in APPENDIX A shall be used to determine compliance. For Max. U value , the values for typical construction assemblies listed in APPENDIX A shall be used to determine compliance. (3) Component U factors for other assemblies (not listed in APPENDIX A) can be determined in accordance with Section A9 of ASHRAE 90.1.2007 Definitions : mass wall : a wall with heat capacity exceeding (1) 7 Btu/ft 2 °F or (2) 5 Btu/ft2 °F, provided that the wall has a material unit weight not greater than 120 lb/ft

3

metal building wall : a wall whose structure consists of metal spanning members supprted by steel structural members (i.e., does not include spandrel glass or metal panels in curtain wall systems) steel framed wall : a wall with a cavity (insulated or otherwise) whose exterior surfaces are separated by steel framing members (i.e. typical steel stud walls & curtain walls system) wood framed & other walls : all other wall types, including wood stud walls.

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FOR GREEN BUILDINGS APPENDIX 3 ASSEMBLY U FACTORS & SHGC FOR UNLABELED VERTICAL FENESTRATION & SKYLIGHTS

(A) FOR VERTICAL FENESTRATION

Frame Type

Glazing Type

Clear Glass U Factor SHGC

Tinted Glass U Factor SHGC

Btu/hr ft2 °f

Btu/hr ft2 °f

All frame types

Single glazing

1.25

0.82

1.25

0.70

Wood, vinyl or fibreglass frames

Double glazing

0.60

0.59

0.60

0.42

Metal & other frames

Double glazing

0.90

0.68

0.90

0.50

(B) FOR SKYLIGHTS To determine the default U factor for unlabeled sloped glazing & skylights without a curb multiply the values in the above APPENDIX 3 by 1.2 To determine the default U factor for unlabeled sloped glazing & skylights on a curb multiply the values in the above APPENDIX 3 by 1.6

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FOR GREEN BUILDINGS APPENDIX 4 MINIMUM EQUIPMENT EFFICIENCY REQUIREMENT AS PER ECBC 2007 Sr. no. Equipment Class

Minimum Minimum COP IPLV

Test Standard

1

Air Cooled Chillers < 530 kW ( = 530 kW (> = 150 TR)

3.05

3.32

ARI 550 / 590 1998

3

Centrifugal Water Cooled Chiller < 530 kW ( = 530 kW & = 150 TR & < 300 TR)

5.8

6.17

ARI 550 / 590 1998

5

Centrifugal Water Cooled Chiller >= 1050 kW (> = 300 TR)

6.3

6.61

ARI 550 / 590 1998

6

Reciprocating Compressors, Water Cooled Chillers all sizes

4.2

5.05

ARI 550 / 590 1998

7

Rotary Screw & Scroll Compressor, Water Cooled Chillers < 530 kW (< 150 TR)

4.7

5.49

ARI 550 / 590 1998

8

Rotary Screw & Scroll Water Cooled Chiller > = 530 kW & = 150 TR & < 300 TR)

5.4

6.17

ARI 550 / 590 1998

9

Rotary Screw & Scroll Water Cooled Chiller >= 1050 kW (> = 300 TR)

5.75

6.43

ARI 550 / 590 1998

Notes:

Heating & Cooling equipment not listed here shall comply with Ashrae 90.1 standard. Unitary Air conditioner shall meet IS 1391 Part 1 Split conditioner shall meet IS 1391 Part 2 Packaged air conditioner shall meet IS 8148

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FOR GREEN BUILDINGS APPENDIX 5 A EQUIPMENT EFFICIENCIES – AIR CONDITIONERS AND CONDENSING UNITS (ASHRAE 90.1.2007)

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FOR GREEN BUILDINGS

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FOR GREEN BUILDINGS APPENDIX 5 B EQUIPMENT EFFICIENCIES WATER CHILLING PACKAGES

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FOR GREEN BUILDINGS APPENDIX 5 C EQUIPMENT EFFICIENCIES – PACKAGE UNITS

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FOR GREEN BUILDINGS

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FOR GREEN BUILDINGS APPENDIX 5 D EQUIPMENT EFFICIENCIES HEAT REJECTION EQUIPMENTS Equipment Type

Rating Conditions

Performance Requireda,b

Test Procedure

Propeller or axial fan cooling towers

95°F entering water 85°F leaving water 75°F wb outdoor air

= 38.2 gpm / hp

CTI ATC-105c & CTI STD-201d

Centrifugal fan cooling towers

95°F entering water 85°F leaving water 75°F wb outdoor air

Air cooled condensers

a b c d

125°F condensing temp. R-22 test fluid 190°F entering gas temp 15°F subcooling 95°F entering db

= 20 gpm / hp

CTI ATC-105 & CTI STD-201

= 1,76,000 Btu/hr.hp

ARI 460

Cooling tower performance is defined as the maximum flow rating of the tower divided by the fan nameplate rated motor power Air cooled condenser performance is defined as the heat rejected from the refrigerant divided by the fan nameplate rated motor power Cooling Technology Institute , Acceptance Test Code for Water Cooling Towers Cooling Technology Institute , Stndard for Certification of Water Cooling Towers Thermal Performance

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FOR GREEN BUILDINGS APPENDIX 5 E EQUIPMENT EFFICIENCIES NON-STANDARD CENTRIFUGAL CHILLERS UPTO 150 TR

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FOR GREEN BUILDINGS APPENDIX 5 F EQUIPMENT EFFICIENCIES NON-STANDARD CENTRIFUGAL CHILLERS FROM 151 TR TO 300 TR

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FOR GREEN BUILDINGS APPENDIX 5 G EQUIPMENT EFFICIENCIES NON-STANDARD CENTRIFUGAL CHILLERS BEYOND 301 TR

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FOR GREEN BUILDINGS APPENDIX 6 PSYCHROMETRIC CHART (FREE COOLING SYSTEM)

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FOR GREEN BUILDINGS APPENDIX 7 FAN POWER LIMITATION PRESSURE DROP ADJUSTMENT

Device Fully ducted return and/or exhaust air systems Return and/or exhaust air flow control devices Exhaust filters, scrubbers, or other exhaust treatment Particulate filtration credit MERV 9 through 12 Particulate filtration credit MERV 13 through 15 Particulate filtration credit MERV 16 and greater and electronically enhanced filter Carbon and other gas phase air cleaners Heat recovery device Evaporative humidifier / cooler in series with another cooling coil Sound attenuation section

Adjustment 0.5" WC 0.5" WC Pressure drop (PD) calculated at fan system design condition 0.5" WC 0.9" WC PD calculated at 2xclean filter PD at fan system design condition Clean filter PD at fan system design condition PD of device at fan system design condition PD of device at fan system design condition 0.15” WC

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FOR GREEN BUILDINGS APPENDIX 8

DUCT INSULATION

Duct Location

Required Insulationa As Per ECBC Required Insulation As Per ASHRAE 90.1 Supply Duct Return Duct Supply Duct Return Duct hr ft2 °f / Btu

Exterior R-7.9 Ventilated Attic R-7.9 Unventilated Attic w/o Roof R-7.9 Insulation Unventilated Attic with Roof R-3.4 Insulation Unconditioned Space R-3.4 Indirectly Conditioned Space No Requirement Buried R-3.4

a

hr ft2 °f / Btu

hr ft2 °f / Btu

hr ft2 °f / Btu

R-3.4 R-3.4

R-6 R-7

R-3.5 R-3.5

R-3.4

R-8

R-3.5

No Requirement

R-3.5

No Requirement

No Requirement No Requirement No Requirement

R-3.5 No Requirement R-3.5

No Requirement No Requirement No Requirement

Insulation R value is measured on a horizontal plane in acoordance with ASTM C518 at a mean temperature of 75°F at installed thk.

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FOR GREEN BUILDINGS APPENDIX 9 PIPING INSULATION

Design Operating Temperature

Required Insulation R-Value hr ft2 °f / Btu

Temp > = 140°F (Heating Systems) 139°F >= Temp > 104°F (Heating Systems) 59°F >= Temp. > 40 (Cooling Systems)

R-4 R-2 R-2

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FOR GREEN BUILDINGS APPENDIX 10 MINIMUM DUCT SEAL LEVEL

Table 1 - Duct Seal Levels Seal Level

Sealing Requirements

A

All transverse joints, longitudinal seams and duct wall penetrations. Pressure sensitive tape shall not be used as primary sealant, unless it has been certified to comply with UL-181A or B by an independent testing laboratory

B

All transverse joints, longitudinal seams. Pressure sensitive tape shall not be used as primary sealant, unless it has been certified to comply with UL-181A or B by an independent testing laboratory Transverse joints only

C

Table 2 - Minimum Duct Seal Level Duct Location

Supply Duct Upto 2" w.c. > 2" w.c. Static Pressure Static Pressure

Outdoor Unconditioned Spaces Conditioned Spaces a

a

A B C

A A B

Return Duct

Exhaust Duct

A B C

C C B

Refer Table-1 for description of seal level

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FOR GREEN BUILDINGS APPENDIX 11 OUTDOOR AIR FLOW CALCULATION

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FOR GREEN BUILDINGS APPENDIX 12 A MINIMUM VENTILATION RATES Default Values Combined Occupant Outdoor Air Density Rate #1000ft2 cfm/person

People Outdoor Air Rate

Area Outdoor Air Rate

cfm/person

cfm/ft2

Cell

5

0.12

25

10

2

Day Room

5

0.06

30

7

1

Guard Stations

5

0.06

15

9

1

Booking/Waiting Educational Facilities

7.5

0.06

50

9

2

Daycare(through age 4)

10

0.18

25

17

2

Classrooms(ages 5-8)

10

0.12

25

15

1

Classrooms(age 9 plus)

10

0.12

35

13

1

Lecture Classroom

7.5

0.06

65

8

1

Lecture hall (fixed seats)

7.5

0.06

150

8

1

Art classroom

10

0.18

20

19

2

Science Laboratories

10

0.18

25

17

-

Wood/Metal shop

10

0.18

20

19

2

Computer Lab

10

0.12

25

15

1

Media Center

10

0.12

25

15

1

Music/Theater/Dance

10

0.06

35

12

1

Multi-use assembly Food and Beverage Service

7.5

0.06

100

8

1

Restaurant Dining Rooms

7.5

0.18

70

10

2

Cafeteria/fast food dining

7.5

0.18

100

9

2

Bars, cocktail lounges General

7.5

0.18

100

9

2

Conference/Meetings

5

0.06

50

6

1

Corridors

Occupancy Category

Notes

Air Class

Correctional Facilities

E

A

-

0.06

Storage Rooms Hotels, Motels, Resorts, Dormitories

0.12

Bedroom/Living Room

5

0.06

10

11

1

Barracks sleeping areas

5

0.06

20

8

1

7.5

0.06

30

10

1

5

0.06

120

6

1

Office Space

5

0.06

5

17

1

Reception areas

5

0.06

30

7

1

Telephone/Data entry

5

0.06

60

6

1

Main entry lobbies Miscellaneous Spaces

5

0.06

10

11

1

Bank vaults/Safe deposit

5

0.06

5

17

2

Computer(not printing)

5

0.06

4

20

1

Pharmacy(prep. Area)

5

0.18

10

23

2

Photo Studios

5

0.12

10

17

1

Lobbies/prefunction Multi-purpose assembly

B

-

1

-

1

Office Buildings

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FOR GREEN BUILDINGS Shipping/Receiving Transportation waiting

-

0.12

B

-

1

7.5

0.06

Warehouses Public Assembly Spaces

-

0.06

100

Auditorium seating area

5

0.06

150

5

1

Places of religious workshop

5

0.06

120

6

1

Courtrooms

5

0.06

70

6

1

Legislative chambers

5

0.06

50

6

1

Libraries

5

0.12

10

17

1

Lobbies

5

0.06

150

5

1

Museums (Childrens)

7.5

0.12

40

11

1

Museums/gallaries Retail

7.5

0.06

40

9

1

Sales (except as below)

7.5

0.12

15

16

2

Mall common areas

7.5

0.06

40

9

1

Barber shop

7.5

0.06

25

10

2

Beauty and nail salons

20

0.12

25

25

2

Pet shops (animal areas)

7.5

0.18

10

26

2

Supermarket

7.5

0.06

8

15

1

Coin-operated Laundries

7.5

0.06

20

11

2

Sports area(play area)

-

0.3

-

Gym, stadium (play area)

-

0.3

30

7.5

0.06

150

-

0.48

Disco/dance floors

20

0.06

100

21

1

Health club/aerobics room

20

0.06

40

22

2

Health club/weight rooms

20

0.06

10

26

2

Bowling alley(seating)

10

0.12

40

13

1

Gambling casinos

7.5

0.18

120

9

1

Game arcaes

7.5

0.18

20

17

1

Stages, studios

10

0.06

70

11

1

B

8

1

-

2

Sports and Entertainment

Spectator areas Swimming (pool and deck)

C

1 2 8

1

-

D

2

GENERAL NOTES FOR TABLE 6.1 1 Related Requirements: The rate in this table are based on all other applicable requirements of this standard being met. 2 Smoking: This table applies to non-smoking areas. Rates for smoking permitted spaces must be determined using other methods. See section 6.2.9 of Ashrae 62.1.2007 for ventilation requirement in smoking areas. 3 Air density: Volumetric airflow rates are based on an air density of 1.2 Kg/m3 which corresponds to dry air at a barometric pressure of 1 atm (101.3 kPa) and an air temperature of 21 deg C. Rates may be adjusted for actual density but such adjustment is not required for compliance with this standard. 4 Default Occupant Density: The default occupant density shall be used when actual occupant density is not known. 5 Default Combined Outdoor Air Rate (per person): This rate is based on default occupant density. 6 Unlisted Occupancies: If the occupancy category for a proposed space or zone is not listed, the requirement for the listed

occupancy category that is most similar in terms of occupant density, activities and building construction shall be used. 7 Healthcare facilities : Rates shall be determined in accordance with Appendix E. ITEM-SPECIFIC NOTES A For high school and college libraries, use values shown for public spaces , library. B Rates may not be sufficient when stored materials include those having potentially harmful emissions. C Rates does not allow for humidity control. Additional ventilation or dehumidification may be required to remove moisture D Rate does not include special exhaust for stage effects, e.g. dry ice vapors, smoke. E No class of air has been established for this occupancy category.

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SHEET 40 OF 48

FOR GREEN BUILDINGS APPENDIX 12 B MINIMUM EXHAUST RATES

Occupancy Category

Exhaust

Notes

Air Class

B

1

Air Rate cfm/ft2 Arenas

0.5

Art Classrooms

0.7

Auto repair rooms

1.5

Barber shops

0.5

2

Beauty & nail Salons

0.6

2

1

2

0.5

2

Darkrooms, Educational Science Labs

1

2

Janitor closets, trash rooms, recycling

1

3

Kitchenettes

0.3

2

Kitchens - Commercial

0.7

2

Locker/dressing rooms

0.25

2

Locker rooms

0.5

2

Cells with toilets Copy, printing rooms

Paint spray booths

2 A

2

--

F

4

Parking garages

0.75

C

2

Pet shops (animal areas)

0.9

Refrigerating machinery rooms

2

--

F

3

50 / 100 per unit

G

2

1

F

3

1.5

F

4

Toilets - private

25 / 50 per unit

E

2

Toilets - public

50 / 70 per unit

D

2

Residential Kitchens Soiled laundry storage rooms Storage rooms, chemical

Woodwork shop/classroom

0.5

2

Notes : A Stands where engines are run shall have exhaust systems that directly connect to the engine exhaust & prevent escape of fumes B When combustion equipment is intended to be used on the playing surface additional dilution ventilation and/or source control shall be provided C Exhaust not required if two or more sides comprise walls that are atleast 50% open to the outside D Rates are per water closet and / or urinal. Provide higher rates where periods of heavy use are expected, such as toilets in theatres, schools , sports facilities etc E Rates is for toilet room intended to be occupied by one person at a time. For continuous system operation during normal hours of use, the lower rate may be used, otherwise use higher rate. F See other applicable standard for exhaust rate. G For continuous system operation, the lower rate may be used. Otherwise use higher rate.

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FOR GREEN BUILDINGS APPENDIX 12 C ZONE AIR DISTRIBUTION EFFECTIVENESS

Air Distribution Configuration

Ez

Ceiling supply of cool air Ceiling supply of warm air & floor return Ceiling supply of warm air 15°F or more above space temperature Floor supply of cool air and ceiling return provided that the 150 fpm supply jet reaches 4.5 ft or more above floor (Note: most underfloor air distribution systems) Floor supply of warm air & ceiling return Make up supply drawn in on the opposite side of the room from the exhaust and / or return Make up supply drawn in near to the exhaust and / or return location

1 1 0.8 1

0.7 0.8

0.5

Ez = Zone air distribution effectiveness

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FOR GREEN BUILDINGS APPENDIX 12 D SYSTEM VENTILATION EFFICIENCY

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SHEET 43 OF 48

FOR GREEN BUILDINGS APPENDIX 13 AIR INTAKE MINIMUM SEPARATION DISTANCE

Object

Minimum Distance (ft)

Significantly contaminated exhaust Noxious or dangerous exhaust (Note 1) Vents, chimneys & flues from combustion appliances & equipments Garage entry, automobile loading area, or drive in queue (Note 2) Truck loading area or dock, bus parking / idling area (Note 2) Driveway, street or parking place (Note 2) Thoughroughfare with high traffic volume Roof, landscaped grade,or other surface directly below intake (Note3) Garbage storage / pickup area Cooling tower intake or basin Cooling tower exhaust

15 30 15 15 25 5 25 1 15 15 25

Note 1 : Laboratory fume hood exhaust air outlets shall be in compliance with NFPA 45 & ANSI std Note 2 : Distance measured to closest place that vehicle exhaust is likely to be located. Note 3 : No min. separation distance applies to surfaces that are sloped more than 45 degrees from horizontal or that are less than 1" wide. Where snow accumulation is expected, distance listed shall be increased by expected avergae snow depth.

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FOR GREEN BUILDINGS APPENDIX 14 LIGHTING POWER DENSITIES Table- 1 Building Area Method

Building area type Automotive faciltiy Convention Center Courthouse Dining : bar lounge/leisure Dining : family Dormitory Exercise Center Gymnasium Health care clinic Hospital Hotel Library Manufacturing facility Motel Motion picture theater Multifamily Museum Office Parking Garage Penitentiary Performing arts theater Police/Fire station Post Office Religious Building Retail School / University Sports arena Town hall Transportation Warehouse Workshop

W/ft2 0.9 1.2 1.2 1.3 1.6 1.0 1.0 1.1 1 1.2 1 1.3 1.3 1 1.2 0.7 1.1 1 0.3 1 1.6 1.0 1.1 1.3 1.5 1.2 1.1 1.1 1 0.8 1.4

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FOR GREEN BUILDINGS

Table- 2 Space by Space Method Common space types

W/ft2

Building specific space types

Office Conference / Meeting / Multipurpose Classroom / Lecture / Training Lobby For Hotel For Performing Art Theater For Motion Picture Theater Audience / Seating area For Gymnasium For Exercise center For Convention center, Penitentiary For Religious building For Sports arena For Performing Art Theater For Motion Picture Theater For Transportation Atrium - First Three floors Atrium - each additional floor Lounge / Recreation For Hospital Dining area For Hotel, Penitentiary For Motel For Bar Lounge / Leisure dining For Family dining Food preperation Laboratory Restrooms Dressing / locker / fitting room Corridor / Transition For Hospital For Manufacturing facility Stairs - Active Active storage For Hospital Inactive storage For Museum Electrical / Mechanical Workshop Sales area (except accent lighting)

1.1 1.3 1.4 1.3 1.1 3.3 1.1 0.9 0.4 0.3 0.7 1.7 0.4 2.6 1.2 0.5 0.6 0.2 1.2 0.8 0.9 1.3 1.2 1.4 2.1 1.2 1.4 0.9 0.6 0.5 1.0 0.5 0.6 0.8 0.9 0.3 0.8 1.5 1.9 1.7

Gym / Exercise center Playing area Exercise area Courthouse/Police Station Courtroom Confinement Cells Judges Chambers Fire Stations Engine room Sleeping quarters Post Office - Sorting area Convention center-Exhibit space Library Card file & cataloging Stacks Reading area Hospital Emergency Recovery Nurses station Exam / Treatment Pharmacy Patient room Operating room Nursery Medical supply Physical Therapy Radiology Laundry-washing Automotive-Service repair Manufacturing Low bay (

Corrugated asbestos, if used for casing, shall be as per IS 459 and of minimum 6 mm thickness. 5.9.3

Plastics Plastics used as materials of construction in cooling towers may be grouped into two classes : (a)

Glass Reinforced Polyesters(GRP) and Fiber Reinforced plastics(FRP) ISSUE R3 TCE FORM NO. 120 R1

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INDUCED DRAFT COOLING TOWERS

GRP or FRP are not suitable when the hot water inlet temperature is above 600C. (b)

Thermoplastics The non-reinforced thermoplastic resins include polyvinyl chloride(PVC) and polypropylene. PVC is not suitable for temperature above 55 0C.

5.9.4

Hardware The choice of materials for cooling tower hardware and fasteners involves the factors of cost, strength, and corrosion resistance. Hardware above and below fan deck level shall be normally provided with hot dip galvanised steel or stainless steel or phosphor bronze or naval brass for fresh water application. Hardware above and below fan deck level shall be of silicon bronze or Super Molybdenum stainless steel for sea water application.

5.9.5

RCC RCC is most commonly used as basin construction material. Large cooling towers shall be constructed of RCC to minimise the risk of fire and for larger structure, higher load carrying capacity.

6. 0

COOLING TOWER COMPONENTS, ACCESSORIES AND BASIN

6.1

CASING Cooling tower casing acts to contain water within tower, provide an air plenum for the fan, and transmit wind loads to the tower frame work. It shall be watertight and shall have corrosion resistant and have fire retardant qualities. It is the enclosure housing the fill (packing) and drift eliminators. It is normally of RCC or treated timber with FRP corrugated panel for large towers, timber with FRP corrugated panel or corrugated asbestos (ACB) for medium size towers and treated timber or FRP for small towers.

6.2

LOUVERS

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INDUCED DRAFT COOLING TOWERS

These are provided on the sides of the tower to equalise the air flow into the fill and to prevent the water drops from falling out. The most utilised louver materials are FRP or corrugated asbestos sheet or treated timber for all types of towers and precast, pre-stressed concrete for RCC towers. 6.3

FILL (PACKING) This is the main component of cooling tower responsible for heat transfer and is of two types i.e. splash type and film type. Advantages of splash fill (a)

It reduces air pressure losses.

(b)

It is not conducive to clogging.

Disadvantages of splash fill (a)

It is very sensitive to inadequate support.

(b)

If proper level is not maintained, there are chances of sagging and channelling of water and air.

Advantage of film fill It requires less space for same amount of cooling than the splash type fill. Disadvantage of film fill The use of film fill shall be avoided in situations where the circulating can become contaminated with debris.

water

Based on type of application, the type of fill shall be decided. For splash type - PVC or FRP or treated timber for all towers and sometimes of RCC for large RCC cooling towers For film type - PVC 6.4

FILL AND TOWER SUPPORTING STRUCTURE

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INDUCED DRAFT COOLING TOWERS

The fill, casing and fan deck are all supported from the basin and the supporting structure shall be of RCC or GRP or CS or SS for large towers, treated timber or GRP or CS or SS for medium and GRP or CS or SS for small towers. 6.5

FAN DECK This is the deck (floor) at the top ( above the fill ) to provide access to the fan and water distribution system. The deck shall be of RCC or treated timber or marine ply timber for large towers and treated timber or marine ply timber for medium and small towers.

6.6

FANS Fans create the required air flow through the tower and are driven by electric motors through gear drives for all towers. The motors shall be suitable for outdoor location. The fan drive shaft shall be of hollow HDGS or stainless steel. The fan blades shall be either of cast aluminium alloy or FRP or GRP. In case of corrosive water, FRP or GRP fan blades shall be used. The gears shall be of enclosed type. The fan blades shall have manually adjustable pitch.

6.7

FAN CYLINDERS AND RECOVERY STACK Recovery stack is provided above fan cylinder to recover kinetic energy of air thereby reducing the power consumption. Suitable air tight access door shall be provided in the cylinder for approach to the fans and for maintenance of gear box and fans etc. Suitable removable rail with tripod and handrail shall be provided for large cooling towers. It is not essential for small towers. Fan cylinder and recovery stack shall be of RCC or FRP or GRP for large towers and treated timber or FRP or GRP for small and medium towers.

6.8

DRIFT ELIMINATORS These are provided at the air outlet from fill, to trap the water particles carried by the air stream and thereby reduce the drift loss. Most widely used drift eliminators material is treated timber or PVC.

6.9

WATER DISTRIBUTION SYSTEM In case of cross-flow tower the hot water is fed to hot water basins or troughs located on the top of tower by means of hot water inlet pipes. From this basin, water flows over the fill by gravity through orifices located in hot water basin.

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INDUCED DRAFT COOLING TOWERS

Counter-flow system normally necessitates the use of a pressure-type system of closed pipe and spray nozzles. The hot water basin is of RCC or marine plywood for large towers and FRP or RCC or timber for medium and small towers. The feed pipes are of hot dip galvanised steel or cast iron. Presently fibre reinforced pipes are used because of low pressure application, light weight, good resistant to both corrosion and erosion. Pre-cast and pre-stressed concrete pipes are also used in RCC tower. The nozzles are of polypropylene or bronze or SS. 6.10

STAIRCASE This is required for providing approach from ground to the fan deck. For large towers, one or two staircases one on either side shall be provided and for medium size tower one staircase is adequate. The staircase is of RCC or treated timber or HDGS for large tower and of treated timber for medium and small towers. For small towers only access ladder of HDG steel or treated timber shall be provided instead of staircase.

6.11

OTHER COMPONENTS AND ACCESSORIES (a)

Fan Deck Illumination This is required for large towers to provide illumination of deck at night.

(b)

Lightning Protection This is also required only for large cooling towers.

(c)

Vibration Limit Switch This is provided to trip the fan motor in case of excessive vibration of fan.

(d)

Vibration Isolators These are provided if considered necessary, to minimise transfer of vibration of rotating parts to fan deck and tower structure.

(e)

Dual Speed Motors

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INDUCED DRAFT COOLING TOWERS

These are provided only in case where the towers are intended for very widely varying duty conditions which is achieved by varying the fan speed

(f)

Fan Guards These are provided on top of fan cylinder or recovery stack of medium and small size towers to prevent birds falling into the stack of standby cell. These are not essential for large towers in view of large size of fan as the fan noise keeps the birds away. The guards are normally HDG steel or SS wire mesh.

(g)

De-icing System This is required to be provided only in place having extreme cold climate which requires passing part of hot water and spraying on to the louvers to prevent the ice formation.

(h)

Hot Water Basin Cover The hot water distribution basin shall be provided with a suitable cover in cross-flow cooling towers to avoid direct sun rays falling on the distribution trough or basin, to minimise algae growth and to prevent choking of distribution nozzles from external falling objects. In case of dusty atmosphere, hot water basin cover shall be provided.

(i)

Low Oil Level Switch This is generally provided to trip the fan in case low oil level in gear box.

(j)

Mechanical Equipment Removal Devices These shall be provided in case of large cooling towers. These shall be provided at fan deck level to remove the mechanical equipment, to lift oil drums and maintenance tools and tackles. This consists of swivel hoist, track and dolley arrangement. The removed component can be dollied to the end of the fan deck, and lowered by means of a properly designed hoisting structure. This structure shall consist of an endwall derrick or an endwall davit depending upon magnitude of load.

(k)

Screens ISSUE R3 TCE FORM NO. 120 R1

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INDUCED DRAFT COOLING TOWERS

Screens shall be provided at the outlet of each basin cell to prevent suspended particles entering to cooling water pumps. Generally, screens of 12 mm mesh opening shall be provided. The material of construction for screen shall be carbon steel with epoxy painting or HDGS in case of fresh water and of SS 316L in case of sea water. Normally, width of screen shall be 1.2 metres. (l)

Stoplogs Insertion plate for water-tight stoplogs shall be provided at the outlet of each basin. Two(2) stoplogs shall be provided to isolate individual basin compartments during maintenance. Material of construction of stoplogs shall be FRP with SS316L frame for sea water and carbon steel with epoxy painting or timber or HDGS for fresh water application. Insertion plate shall be of SS316L for sea water and carbon steel with epoxy painting for fresh water.

(m) 6.12

Monorail with hoist shall be provided for lifting screens and stoplogs.

COLD WATER BASIN This holds the re-cooled water falling through the fill and provides an outlet for the cooled water. The basin is of RCC for large and medium size towers and of steel with epoxy painting or FRP or steel for smaller towers. Basin depth may be decided based on following considerations: (a)

Flooded suction and NPSH required for cooling water pumps, if horizontal pumps are used and submergence and pump dimensions if vertical pumps are used.

(b)

Cooling tower should be able to supply cooling water within acceptable temperature limits under variable cooling loads during normal operation.

(c)

Cooling tower should be able to supply cooling water within acceptable temperature limits during start-up, shut-down and emergency conditions, if so specified.

(d)

Basin should be able to hold back-flow during pump shut-down and cooling water shall not overflow into drains. This is particularly important if treated

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water is used as cooling water make-up or water is scarce and wastage is to be minimized. (e)

Basin should have enough reserve capacity during interruption of make-up water supply.

(f)

If basin is to be used as a reservoir for fire water supply, take inputs from Basic Study for Fire Protection System. For underground basins, soil conditions and ground water table etc. may be checked with civil department while deciding the depth of the basin.

(g)

The large and medium towers shall be supported on RCC foundation whereas small cooling tower shall be supported on either concrete pedestals or steel sections. The basin for large towers may be completely underground whereas the basins for medium size and small towers may be above ground level, if required, to provide flooded suction for the pumps. The cold water outlet from basin shall be through channel for large cooling towers and through pipes for medium and small towers. Normally, the cold water basin may be partitioned cell wise to facilitate cleaning during individual cell maintenance. Outlet of each partition shall be connected to common sump. In case of large cooling towers with more than 5 cells, the cold water basin may be partitioned into two compartments along the length of tower and cleaning of one compartment of cold water basin may be carried out by shutting down half part of individual cells of entire cooling tower. 7.0

EVALUATION AND PENALTY FACTORS

7.1

FOR PUMPING HEAD Increase in cooling tower height i.e. height from basin curb to centre line of riser at fan deck increases the cooling water pump head For cooling water circulation pumps SxQxH BKW = 367.2 x η where S = specific gravity of water ISSUE R3 TCE FORM NO. 120 R1

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INDUCED DRAFT COOLING TOWERS

Q = Capacity in M3/Hr H = Differential Head in M η = Efficiency of pumps For every metre excess head, there is rise in pump power consumption. For every rise in excess BKW, the penalty factor (P) shall be calculated as follows: P =

where N =

Energy charge in Rs/KwHr x number of operating hours per year x N x differential power Capitalisation Factor ( 1 + r/100)n - 1

= r/100 x (1 + r/100)n r n P 7.2

= = =

Rate of interest in % number of years Penalty in Rs.

FOR FAN POWER CONSUMPTION For every KW differential of excess fan power consumption, penalty factor P shall be calculated as shown in para 7.1

8.0

REFERENCES References for design and construction features, thermodynamic aspects and testing requirements of cooling towers are given below : (a)

The Industrial Cooling Tower by Mckelvey and Brooke

(b)

BS 4485 Part 1 Part 2 Part 3 Part 4 -

(c)

Cooling tower performance curves published by Cooling Tower Institute (CTI)

Specification for Water Cooling Towers Glossary of Terms Methods of Performance Testing Thermal and Functional Design Structural Design of Cooling Towers

ISSUE R3 TCE FORM NO. 120 R1

TCE.M6-ME-127-214

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR

SECTION: WRITE-UP

SHEET 17 OF 20

INDUCED DRAFT COOLING TOWERS

(d)

CTI acceptance test procedure as per CTI code ATC 105 for Water Cooling Towers

(e)

Cooling Tower Manual

(f)

CTI Bulletin WMS-112, " Pressure Preservatives Treatment of Lumber for Industrial water"

ISSUE R3 TCE FORM NO. 120 R1

TCE.M6-ME-613-215

TCE CONSULTING ENGINEERS LIMITED

SECTION: TITLE

DESIGN GUIDE FOR HORIZONTAL CIRCULATING WATER PUMPSETS

SHEET i OF iii

DESIGN GUIDE FOR HORIZONTAL CIRCULATING WATER PUMPSETS

FILE NAMES: M6ME215R4.DOC AND M6ME215R4.DWG

REV. NO.

R1

R2

R3

R4 ISSUE

INITIALS

SIGN.

INITIALS

SIGN.

INITIALS

SIGN.

INITIALS

PPD. BY

DGR

Sd/-

SCP

Sd/-

SHN

Sd/-

VB

CHD. BY

RKJ

Sd/-

MPB

Sd/-

RKJ

Sd/-

PDG

APD. BY

KG

Sd/-

SCM/RL

Sd/-/Sd/-

SCM/RL

Sd/-/Sd/-

RL

SIGN.

R4

DATE

02.05.1994

06.08.1999

17.08.2002

28.10.2003 TCE FORM NO. 020R2

TCE.M6-ME-613-215

TCE CONSULTING ENGINEERS LIMITED

SECTION: CONTENTS

DESIGN GUIDE FOR HORIZONTAL CIRCULATING WATER PUMPSETS

SHEET ii OF iii

CONTENTS SL. NO.

TITLE

SH. NO.

1.0

SCOPE

1

2.0

INPUT DATA REQUIRED

1

3.0

APPLICATION

1

4.0

SELECTION OF PARAMETERS

1

5.0

TYPE OF PUMP

3

6.0

CONSTRUCTION FEATURES

3

7.0

MATERIALS OF CONSTRUCTION

4

8.0

PERFORMANCE CHARACTERISTICS

5

9.0

LAYOUT CONSIDERATIONS

5

10.0

MOTORS

5

11.0

PERFORMANCE TEST

5

12.0

NOISE AND VIBRATION

5

13.0

CROSS REFERENCES

6

APPENDICES 1.

COOLING WATER REQUIREMENTS

7

2.

MATERIALS OF CONSTRUCTION

8

3.

PRESSURE IMPELLER

4.

DOUBLE FLAT RING TYPE WEARING RING

10

5.

CONVENTIONAL STUFFING BOX

11

6.

LAYOUT OF SUCTION PIPING

12

ACTING

ON

A

DOUBLE

SUCTION

9

ISSUE R4 TCE FORM NO. 120R1

TCE.M6-ME-613215

TCE CONSULTING ENGINEERS LIMITED

SECTION: REV. STATUS

DESIGN GUIDE FOR HORIZONTAL CIRCULATING WATER PUMPSETS

SHEET iii OF iii

REVISION STATUS REV. NO.

DATE

DESCRIPTION

R0

24.07.1986

--

R1

02.05.1994

Contents Sheet, Write-up Sheets 3, 4, 9, 11, 12 and Table II revised.

R2

06.08.1999

Overall Revision

R3

17.08.2002

Clauses 3.0, 4.1, 4.2.1 and 10.0 revised. Clause 4.3.2 added.

R4

28.10.2003

Appendix 2 revised.

ISSUE R4 TCE FORM NO. 120R1

TCE.M6-ME-613-215

1.0

TCE CONSULTING ENGINEERS LIMITED

SECTION: WRITE-UP

DESIGN GUIDE FOR HORIZONTAL CIRCULATING WATER PUMPSETS

SHEET 1 OF 12

SCOPE This design guide deals with selection of parameters, types, construction features, materials of construction, performance characteristics, tests etc. for horizontal circulating water pumpsets of cooling system for power plants.

2.0

INPUT DATA REQUIRED The following data shall be obtained for the selection of the proper pumpset for the intended service:

3.0

(a)

Cooling water requirement of the plant, with complete break-up

(b)

Normal, maximum and minimum water level in the sump

(c)

Minimum and maximum water temperature

(d)

Analysis of water

(e)

Information regarding screening of water ahead of the pumps by means of trash racks/screens.

(f)

Chlorine dosage and maximum residual chlorine

APPLICATION Normally, it is the practice to keep the plant grade level above the highest flood level to avoid flooding of the power station. Keeping in mind the above, usually it will not be possible to go in for horizontal pumps for once through cooling system due to the high suction lift because of variation in water level at the source. Hence, possibility of using horizontal pumps for once through system is very remote. In cooling tower installations sometimes it becomes necessary to locate the cooling tower basin and cold water channels above the grade level due to the soil conditions at site. It is preferable to go in for horizontal pumps for such installations. It is recommended to finalise the choice in this case, after cost comparison with the option of below ground basin with vertical turbine pumps.

4.0

SELECTION OF PARAMETERS The three important parameters that are to be decided for the pumps are:

ISSUE R4 TCE FORM NO. 120R1

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SECTION: WRITE-UP

DESIGN GUIDE FOR HORIZONTAL CIRCULATING WATER PUMPSETS

SHEET 2 OF 12

(a) (b)

Capacity Total head

(c)

Speed

4.1

CAPACITY

4.1.1

The capacity of the pump depends on the total quantity of water required for the circulating water system and the number of pumps to be provided for the system. The number of pumps provided could be any of the following alternatives: (a)

Two numbers each of 50% capacity

(b)

Three numbers each of 33.3% capacity

(c)

Four numbers each of 25% capacity

4.1.2

The approximate cooling water requirements for different sizes of generating units are given Appendix-1.

4.1.3

The quantity of standby pumps shall be finalised after discussion with client. Typical guidelines could be:

4.2

(a)

Nil or minimum standby pumps for fresh water application.

(b)

One number standby for (a) and (b) above for sea water application.

(c)

For configuration (c) above for seawater application, quantity could be decided after doing economic analysis of providing standby pump vis-à-vis loss of generation.

TOTAL HEAD Total head comprises static head and friction losses.

4.2.1

Static Head The static head shall be either the difference in elevation between the water level in the circulating water sump and the maximum water level or discharge pipe elevation whichever is higher. For efficient operation of the circulating water pumpsets the normal water level in the sump shall be considered for computation of the static head. Minimum water level in sump could be considered in case a prolonged duration of water level being at minimum level is anticipated. ISSUE R4 TCE FORM NO. 120R1

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4.2.2

TCE CONSULTING ENGINEERS LIMITED

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DESIGN GUIDE FOR HORIZONTAL CIRCULATING WATER PUMPSETS

SHEET 3 OF 12

Friction Head Friction head is the head necessary to overcome the friction losses in the heat exchanger, piping, valves and fittings for the system in which the pump operates. The friction head depends on several factors like velocity of water, type of fittings, inside surface roughness of the pipeline, length of the pipeline involved in the system etc. The method of calculation of friction head for a given system is explained in the Design Guide No. TCE.M6-ME-613-212 “Guide for Calculation of Hydraulic Losses for Water in Pipes, Fittings and Valves”.

4.3

SPEED

4.3.1

Selection of speed is governed by following considerations: (a)

Type of driver contemplated for the unit

(b)

Higher specific speed results in a smaller pump and cheaper drivers.

4.3.2

The pump parameters shall be achievable at a speed corresponding to normally available frequency.

5.0

TYPE OF PUMP The pumps could be classified according to the orientation of the suction and discharge nozzles or could be classified according to the design of the casing. Side suction and side discharge (horizontal split casing pumps) are generally used. In horizontal split casing pumps the casing is divided by a horizontal plane through the shaft centre line or axis. Since both suction and discharge nozzles are in the lower half of the casing the upper half of the casing could be removed for inspection of the interior without disturbing the bearings or the piping. The benefit of this type is that the mechanical design of casing is structurally stable, since the impeller is supported by bearings on either side of the impeller.

6.0

CONSTRUCTION FEATURES

6.1

ENCLOSED IMPELLER Only enclosed impellers of double suction type are used for these pumps. The impeller incorporates shroud that totally encloses the impeller waterways from the suction eye to the periphery. This design prevents the leakage of water between the impeller and ISSUE R4 TCE FORM NO. 120R1

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SHEET 4 OF 12

its side plates. However, a running joint must be provided between the impeller and the casing to separate the discharge and suction chambers of the pumps. The thrust developed for a closed impeller is minimum. 6.2

DOUBLE SUCTION IMPELLER In a double suction impeller theoretically a hydraulic axial balance exists with the pressure on one side of the impeller equal to and counter balancing the pressure on the other side of the impeller. However, in practice it may not be possible to achieve this hydraulic balance. The forces acting on a double suction impeller is shown in Appendix-3.

6.3

WEARING RINGS Wearing rings provide an easily and economically renewable leakage joint between impeller and casing. The material of the wearing ring shall be softer than the impeller material so that impeller does not get worn out. Some times pumps are provided with two wearing rings, one on the casing and the other on the impeller. A double flat-ring construction type of wearing ring is shown in Appendix-4.

6.4

SHAFT COUPLING The pump shaft is to be connected to the motor shaft by means of flexible/spacer coupling.

6.5

SHAFT SLEEVE The shaft sleeve to be provided to protect shaft from direct wear.

6.6

STUFFING BOX If the pump works with suction lift and the pressure at the interior of stuffing box end is below atmospheric, stuffing box prevents air leakage into the pump. If the pressure inside the pump is above atmospheric the stuffing box prevents the liquid being pumped from leaking out. A conventional stuffing box used in a horizontal pump is shown in Appendix-5. Mechanical seal instead of gland packing also can be provided to prevent leakage.

6.7

LUBRICATION AND COOLING OF PUMPS The stuffing box gland/mechanical seal shall be supplied water for lubrication/cooling/sealing of the stuffing box packing/mechanical seal. The water for lubrication/cooling/sealing shall be taken from the pump discharge itself. When pump ISSUE R4 TCE FORM NO. 120R1

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DESIGN GUIDE FOR HORIZONTAL CIRCULATING WATER PUMPSETS

SHEET 5 OF 12

is handling seawater, filtered seawater should be used. 7.0

MATERIALS OF CONSTRUCTION The materials of construction of the various components of the pump shall be selected based on the type of water being handled. The materials shall be so selected that corrosion of the parts is prevented which in turn reduces the maintenance of the pump. The materials to be adopted for various components of the pump are indicated in the Appendix-2.

8.0

PEFORMANCE CHARACTERISTICS The pump at constant speed delivers any capacity from zero to maximum value depending on the size, design and suction condition of the pump. The pump shall have stable rising (H Vs Q) characteristic curve from run out flow to no flow condition. The maximum head developed by the pump shall be at shut off condition. The pumps shall operate at the best efficiency point. At least 1 MLC difference shall be maintained between (NPSHa – NPSHr) throughout NPSH curve.

9.0

LAYOUT CONSIDERATIONS The sump shall be so designed that the velocity of water in the sump is sufficiently low not to cause any turbulence near the pump. The sump dimensions shall be based on the recommendations of Hydraulic Institutes Standard. Layout of suction piping with tentative dimensions is given in Appendix-6 (also refer design guide TCE.M6-ME613-213 ”Design Guide for Pump Sump”).

10.0

MOTORS Normally, circulating water pumps require 6.6 kV or 3.3 kV, 3 phase, 50 Hz, motors. The selection of the proper motor rating is a very important factor in satisfactory operation of the pump sets. The motor rating shall be at least 116% of the power required by the pump at duty point. The enclosure for the motors shall be Totally Enclosed Fan Cooled (TEFC), Circulating Air Closed Air (CACA) or Circulating Air Closed Water (CACW). CACA motors though costlier than CACW motors, are normally preferred as it eliminates the water circulation arrangement for motor.

11.0

PERFORMANCE TEST Performance test of prototype pump shall be conducted at manufacturer’s works as per BS: 5316, Part 2/ISO 3555 to find out whether the pumps meet the guaranteed values of capacity, head, efficiency and power input.

ISSUE R4 TCE FORM NO. 120R1

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12.0

TCE CONSULTING ENGINEERS LIMITED

SECTION: WRITE-UP

DESIGN GUIDE FOR HORIZONTAL CIRCULATING WATER PUMPSETS

SHEET 6 OF 12

NOISE AND VIBRATION After installation of pumpsets the following parameters are to be checked:

13.0

(a)

The noise level of the pumpset shall not exceed 85 dBA measured at a distance of 1.86 m from the outline of the equipment.

(b)

The velocity of vibration shall not exceed 4.5 mm/sec (Refer ISO-108161/BS: 7854 Part-I).

CROSS REFERENCES This document makes reference to the following standard documents: DOCUMENT NO.

DOCUMENT TITLE

TCE.M6-ME-613-212

Guide for Calculation of Hydraulic losses for Water in Pipes, Fittings and Valves

TCE.M6-ME-613-213

Design Guide for Pump Sump

ISSUE R4 TCE FORM NO. 120R1

TCE.M6-ME-613-215

TCE CONSULTING ENGINEERS LIMITED

SECTION: APPENDIX

DESIGN GUIDE FOR HORIZONTAL CIRCULATING WATER PUMPSETS

SHEET 7 OF 12

APPENDIX - 1 COOLING WATER REQUIREMENTS MW/UNIT

TOTAL COOLING WATER REQUIRED (M3/Hr)

50

10,000

110

17,000 to 18,000

200

32,000

500

66,000

ISSUE R4 TCE FORM NO. 120R1

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SECTION: APPENDIX

DESIGN GUIDE FOR HORIZONTAL CIRCULATING WATER PUMPSETS

SHEET 8 OF 12

APPENDIX-2 MATERIALS OF CONSTRUCTION MATERIAL OF CONSTRUCTION SL. NO.

COMPONENT

SEA WATER SERVICE ALTERNATIVE-1

ALTERNATIVE-2

FRESH WATER SERVICE

1.

Casing

Ni-resist as per BS 3468 GR S2W

2.

Impeller

Stainless steel as per Stainless steel as per Bronze as per IS 318 ASTM A351 GR ASTM A351 GR GR LTB2 OR Cr-Ni CF3M OR CF8M CF3M OR CF8M alloy as per ASTM A743 GR CF-8

3.

Shaft

Cr-Ni alloy as per Cr-Ni alloy as per Carbon steel as per IS ASTM A743 GR CK- ASTM A743 GR CK- 1570 40C8 OR ASTM 20 20 OR Cr-Ni-Mo alloy A107 GR 1040 as per ASTM A743 GR CF-8M

4.

Shaft Sleeves

Cr-Ni alloy as per Cr-Ni alloy as per Bronze as per IS 318 ASTM A743 GR CK- ASTM A743 GR CK- GR LTB2 20 20 OR Cr-Ni-Mo alloy as per ASTM A743 GR CF-8M

5.

Wearing rings/ Casing rings

Ni-resist as per BS 3468 GR S2W

Ni-resist as per ASTM A436 Type 2

Leaded bronze

6.

Gland

Ni-resist as per BS 3468 GR S2W

Ni-resist as per ASTM A436 Type 2

Bronze as per IS 318 GR LTB2

7.

Lantern ring

Ni-resist as per BS 3468 GR S2W

Ni-resist as per ASTM A436 Type 2

Bronze as per IS 318 GR LTB2

8.

Packing

PTFE

PTFE

Graphited asbestos

9.

Base plate

Carbon painted

10.

Hardware in contact with water

Stainless steel as per Stainless steel as per Stainless steel as per ASTM A276 Type ASTM A276 Type ASTM A276 Type 410 316L 316L

steel

Ni-resist as per ASTM Cast iron as per IS 210 A436 Type 2 OR GR FG 220 OR ASTM Stainless steel as per A48 CL 35 ASTM A351 GR CF3M OR CF8M

epoxy Carbon painted

steel

epoxy Fabricated steel as per IS 2062

ISSUE R4 TCE FORM NO.120R1

TCE.M6-ME-613-215

TCE CONSULTING ENGINEERS LIMITED

SECTION: APPENDIX

DESIGN GUIDE FOR HORIZONTAL CIRCULATING WATER PUMPSETS

SHEET 9 OF 12

APPENDIX-2 MATERIALS OF CONSTRUCTION MATERIAL OF CONSTRUCTION SL. NO.

COMPONENT

SEA WATER SERVICE ALTERNATIVE-1

11.

Hardware not in contact with water

ALTERNATIVE-2

FRESH WATER SERVICE

Carbon steel as per IS Carbon steel as per IS Carbon steel as per IS 1367-Class 6.6 1367-Class 6.6 1367-Class 6.6

ISSUE R4 TCE FORM NO.120R1

TCE.M6-ME-610-217

TCE CONSULTING ENGINEERS LIMITED

SECTION: TITLE

DESIGN GUIDE FOR TOTAL HEAD AND NPSH CALCULATIONS FOR CENTRIFUGAL PUMPS

SHEET i OF iii

DESIGN GUIDE FOR TOTAL HEAD AND NPSH CALCULATIONS FOR CENTRIFUGAL PUMPS

FILE NAME: M6ME217R4.DOC REV. NO.

R1

R2

R3

R4 ISSUE

INITIALS

SIGN.

INITIALS

SIGN.

INITIALS

SIGN.

INITIALS

PPD. BY

DVL

Sd/-

MMK

Sd/-

MMK

Sd/-

SDP

CHD. BY

OKM

Sd/-

TSR

Sd/-

TSR

Sd/-

TSR

APD. BY

KG

Sd/-

RL

Sd/-

RL

Sd/-

RL

SIGN.

R4

DATE

29.11.984

17.01.1998

02.01.2001

02.02.2004 TCE FORM NO. 020 R2

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-610-217

DESIGN GUIDE FOR TOTAL HEAD AND NPSH CALCULATIONS FOR CENTRIFUGAL PUMPS

SECTION: CONTENTS

SHEET ii OF iii

CONTENTS SL. NO.

TITLE

SH. NO.

1.0

SCOPE

1

2.0

TYPICAL PUMP SKETCH

1

3.0

TOTAL HEAD CALCULATIONS

1

4.0

NPSH CALCULATIONS

2

ISSUE R4 TCE FORM NO. 120 R1

TCE.M6-ME-610-217

TCE CONSULTING ENGINEERS LIMITED

SECTION: REV. STATUS

DESIGN GUIDE FOR TOTAL HEAD AND NPSH CALCULATIONS FOR CENTRIFUGAL PUMPS

SHEET iii OF iii

REVISION STATUS REV. NO.

DATE

DESCRIPTION

R0

10.09.1981

--

R1

29.11.1984

--

R2

17.01.1998

Reformatted in MS Word.

R3

02.01.2001

Generally revised.

R4

02.02.2004

Reformatted.

ISSUE R4 TCE FORM NO. 120 R1

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1.0

SECTION: WRITE-UP

DESIGN GUIDE FOR TOTAL HEAD AND NPSH CALCULATIONS FOR CENTRIFUGAL PUMPS

SHEET 1 OF 2

SCOPE This document gives guidelines for calculating total head and Net Positive Suction Head (NPSH) for centrifugal pumps.

2.0

TYPICAL PUMP SKETCH EL.Ed

Pd

EL.En EL.Es

Ps

PSSSSS DISCHARGE VESSEL

SUCTION VESSEL

C

EL.Ep

PUMP

3.0

TOTAL HEAD CALCULATIONS

SL. NO.

ITEM

SYMBOL

UNIT

3.1

Elevation of water level in suction vessel

Es

M

3.2

Elevation of water level in discharge vessel

Ed

M

3.3

Elevation of centre-line nozzle on discharge vessel

of

En

M

3.4

Elevation of centre-line of pump

Ep

M

3.5

Static suction head

hs

MLC

3.6

Friction losses in suction piping and entry losses

hfs

MLC

3.7

Gauge pressure in suction vessel

Ps

Kgf/ cm2g

3.8

Specific gravity of the liquid handled

S

FORMULA

VALUE

Es - Ep

ISSUE R4 TCE FORM NO. 120 R1

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SECTION: WRITE-UP

DESIGN GUIDE FOR TOTAL HEAD AND NPSH CALCULATIONS FOR CENTRIFUGAL PUMPS

SL. NO.

ITEM

SHEET 2 OF 2

SYMBOL

UNIT

FORMULA

3.9

Suction head

Hs

MLC

hs - hfs + (Ps × 10 )/S

3.10

Static discharge head

hd

MLC

(Greater of Ed or En) Ep

3.11

Friction losses in discharge piping, equipment and the exit losses

hfd

MLC

3.12

Gauge pressure in discharge vessel

Pd

Kgf/ cm2 g

3.13

Delivery head

Hd

MLC

hd + hfd + (Pd × 10)/S

3.14

Total head

H

MLC

Hd - Hs

FORMULA

4.0

VALUE

NPSH CALCULATIONS

SL.NO.

ITEM

SYMBOL

UNIT

4.1

Absolute vapour pressure at the temperature at which liquid is handled

Pv

Kgf/ cm2

4.2

Absolute atmospheric pressure

Pa

Kgf/ cm2

4.3

Net Positive Suction Head Available (NPSH A)

NPSH

VALUE

MLC (Pa × 10)/S + Hs (Pv × 10 )/S

NOTES 1.

All elevations shall be considered from common datum.

2.

Elevations mentioned correspond to normal conditions. However, these shall be crosschecked with the extreme conditions.

3.

Ps and Pd are positive if the pressure is above atmospheric pressure and negative if the pressure is below atmospheric pressure.

4.

Hs and hs may either be positive or negative. ISSUE R4 TCE FORM NO. 120 R1

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SECTION: TITLE

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SHEET i OF iii

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

REV. NO.

R0

R1

R2

INITIALS

SIGN.

INITIALS

SIGN.

INITIALS

PPD. BY

PRJ

Sd/-

HRK

Sd/-

HRK

CHD. BY

DJ

Sd/-

PRJ

Sd/-

PRJ

APD. BY

JSK

Sd/-

RL

Sd/-

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DATE

22.12.1988

27.12.1999

ISSUE SIGN.

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01.12.2002 TCE FORM NO. 020R2

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DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

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REV. NO.

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HRK

Sd/-

HRK

CHD. BY

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Sd/-

PRJ

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PRJ

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Sd/-

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Sd/-

RL

DATE

22.12.1988

27.12.1999

ISSUE SIGN.

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01.12.2002 TCE FORM NO. 020R2

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SECTION: CONTENTS SHEET ii OF iii

CONTENTS

SL. NO.

TITLE

SH.NO.

1.0

SCOPE

1

2.0

INPUT DATA

1

3.0

COMPONENTS AIR-CONDITIONING LOAD

2

4.0

ESTIMATION OF AIR-CONDITIONING LOAD

5

5.0

APPARATUS DEW POINT

27

6.0

DEHUMIDIFIED AIR FLOW RATE

28

7.0

REFERENCES

28

APPENDICES 1.

TYPICAL AIR-CONDITIONING PSYCHROMETRIC PROCESS

29

2.

SAMPLE AIR-CONDITIONING LOAD CALCULATIONS

30

3.

ASHRAE PSYCHROMETRIC CHART NO. 1 FOR NORMAL TEMPERATURES AT SEA LEVEL

41

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SECTION: REV.STATUS SHEET iii OF iii

REVISION STATUS

REV. NO.

DATE

DESCRIPTION

R0

88.12.22

----------------

R1

99.12.27

Completely revised.

R2

02.12.01

Generally revised.

ISSUE R2 TCE FORM NO. 120R1

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1.0

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SECTION: WRITE-UP SHEET 1 OF 41

SCOPE This document outlines the procedure for performing air-conditioning cooling load, heating load and dehumidified air flow rate calculations.

2.0

INPUT DATA Following input data shall be collected from various sources like client, process collaborator, other groups in the mechanical department, other departments, bidders and vendors.

2.1

Location of the site and the plot plan

2.2

Latitude and altitude of the site

2.3

Names of areas to be air-conditioned

2.4

Equipment layout drawings of the areas to be air-conditioned

2.5

Architectural drawings and civil Reinforced Cement Concrete (RCC) or structural steel drawings of the areas to be air-conditioned

2.6

Outside design conditions like Dry Bulb Temperature (DBT), Wet Bulb Temperature (WBT) and Relative Humidity (RH) for summer, monsoon and winter and daily temperature range

2.7

Inside design conditions and internal loads for each of the areas to be airconditioned (a)

DBT range

(b) RH range (c)

Lighting loads

(d) Sensible and latent equipment heat loads of heat sources within and adjacent to the air-conditioned areas with specific periodic loading patterns, if any

2.8

(e)

Occupancy

(f)

Fresh air and exhaust air requirements, if any

(g)

Filtration and cleanliness requirements, if any

(h)

Pressurisation requirements, if any

(i)

Air flow pattern requirements, if any

(j)

Hazardous area classification

Duration of air-conditioning for each of the areas to be air-conditioned ISSUE R2 TCE FORM NO. 120R1

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SECTION: WRITE-UP SHEET 2 OF 41

2.9

Availability of cooling water, make-up water, quality of water, inlet and outlet pressures and temperatures

2.10

Specific requirements like future expansion plans and provisions required to be incorporated at present Some of the input data indicated above may not be readily available at the time of performing these calculations. Appendices 1 to 4 of TCE.M1-ME-811-301 give guidelines for assuming such data.

3.0

COMPONENTS OF AIR-CONDITIONING LOAD

3.1

STRUCTURAL HEAT GAIN

3.1.1

Sensible Heat Gain (a)

External Walls This component of heat gain is by virtue of the Equivalent Temperature Difference (ETD) across walls, ceiling, exposed roof or floor RCC slabs and door or window glass. The ETD is the DBT difference across the wall and the outdoor. Additional corrections in the temperature differential are due to conduction across the surface, convection and incident radiation on the surface, month of the year, time of the day, storage effect due to the heat capacity of the structure, the latitude and the orientation of the building or structure with respect to the true north. The estimation of this component of heat gain requires the estimation or establishing of the following parameters for each of the surfaces of the structure which are exposed to the outside:

(b)

(i)

Heat transfer area

(ii)

Heat transfer coefficient

(iii)

ETD

Internal Walls (Partition Walls) This component of heat gain is by virtue of the temperature difference across the internal walls of the building, i.e. the areas of the building adjacent to the area to be air-conditioned. The estimation of this component of heat gain requires the estimation or establishing of the following parameters for each of the surfaces of the structure which are not exposed to the outside: (i)

Heat transfer area

(ii)

Heat transfer coefficient ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS (iii)

3.1.2

SECTION: WRITE-UP SHEET 3 OF 41

Temperature difference (∆T)

Latent Heat Gain The structural heat gain also includes the latent heat load due to the moisture transmission across the walls, roof or ceiling and floor slabs of the structure or by ingress with the air that enters the room through door openings, fixed wall openings for conveyers or open windows. This load component is especially significant for areas having inside design RH values of 45% and lower. The moisture transmission load is a function of the differential vapour across the structure.

3.2

FENESTRATION HEAT GAIN This component of heat gain considers the radiation entering the air-conditioned area through exposed glass surfaces. The conduction and convection components of fenestration heat gain have been described in para 3.1.1. This component is not normally applicable to internal walls. The estimation of this component requires the estimation or establishing of the following parameters:

3.3

(a)

Glass area

(b)

Glass transmittance factor

(c)

Glass Solar Heat Gain Factor (SHGF)

INFILTRATION HEAT GAIN This component is due to ambient air leaking into the air-conditioned area through gaps between door or window frames and their shutters, cracks in the walls etc. The estimation of this component requires the estimation or establishing of the following parameters:

3.4

(a)

Area of gaps between door and window frames, cracks in walls etc.

(b)

Air pressure difference across the gap or crack

INTERNAL HEAT GAIN This component is due to heat dissipation by the equipment, people and lighting. The estimation of this component requires the estimation or establishing of the following parameters: (a)

Sensible and latent heat dissipation loads from equipment, the locations and quantities of the equipment

(b)

Number of people occupying each of the areas and the activities performed by these people ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

(c)

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SECTION: WRITE-UP SHEET 4 OF 41

Number, type (fluorescent or incandescent) of lighting fixtures in each of the areas and their rating in watts. In some cases special fixtures are provided which also serve as return air inlets. Such fixtures (Troufers) do not reduce the total heat dissipation from the lighting fixture but reduce the lighting heat dissipation into the room. The reduction in the lighting heat dissipation into the room is instead transferred to the return air, reducing the room sensible heat gain and consequently the room dehumidified air flow rate. This heat dissipation load, within the room and in the return air, shall be furnished by the electrical department from approved vendor drawings or catalogues.

The diversity factor for the heat dissipation from each of the above shall be considered. 3.5

FAN AND DUCT HEAT GAIN This component is due to the supply air fan power consumption and the heat gain to the supply and return air duct. In addition to the above, a notional heat gain is considered for the leakage from the supply air and the return air. Heat gain from return air fan, if provided, shall be added to the return air.

3.6

FRESH AIR HEAT GAIN This component is due to the fresh air introduced into the air-conditioned area to maintain indoor air quality or pressurisation. The estimation of this component requires the estimation or establishing of the following parameters:

3.7

(a)

Quantity of fresh air

(b)

Cooling coil bypass factor

RETURN AIR HEAT GAIN The heat gain to the return air, after collection from the area to be air-conditioned may form a significant proportion of the total cooling load. This cooling load does not form part of the room cooling load and therefore has no contribution to the dehumidified air flow rate calculation. The return air heat gain shall be estimated for the heat gain through the return air duct, which may be sheet metal or masonry. The heat gain to return air, when collected above the false ceiling, from exposed roof or ceiling is theoretically not part of the room load and therefore should not be considered as the room heat gain which is considered for the estimation of the dehumidified air flow rate. However, to simplify the procedure for the airconditioning load calculation, this load is considered as part of the room load. This will give about 2 to 3% higher dehumidified air flow rate which is more conservative and acceptable. This return air heat gain above the false ceiling can, if so desired by ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SECTION: WRITE-UP SHEET 5 OF 41

the designer, be estimated by an iterative procedure and considered for the dehumidified air flow rate estimation. The parameters which are to be established are : (a)

Overall heat transfer coefficient across the return air duct

(b)

Area of the return air duct

(c)

ETD for the return air duct

ETD shall be estimated for return air routed through masonry or exposed insulated sheet metal ducts, which are exposed to the ambient. In case the ducts are located indoors, the DBT difference shall be the temperature difference across the return air duct wall. 4.0

ESTIMATION OF AIR-CONDITIONING LOAD

4.1

STRUCTURAL HEAT GAIN

4.1.1

Sensible Heat Gain (a)

Heat Transfer Area The areas of the heat transfer surfaces of each the air-conditioned areas shall be estimated from the architectural drawings. The heat transfer areas include the following: (i)

Wall face areas. These include walls exposed to the ambient (exposed walls) and walls which are internal to the building.

(ii)

Floor, ceiling or exposed roof RCC slab

(iii)

Glass areas. These are for doors or windows. These, like walls, could also be exposed or internal.

The orientation of the respective exposed walls and glass surfaces, above and below the false ceiling, shall be noted with respect to the true north. The wall and slab area shall be estimated considering the grouping of the various areas, for the purpose of combining the areas under one unit or under multiple units, each being operated separately. For example, in case two adjacent areas are being air-conditioned simultaneously and the inside design temperatures are similar, there will be no heat transfer between these areas. The common glass, wall or RCC slab between these areas shall not be considered for heat transfer area calculations. In case the areas have different hours of operation of the air-conditioning systems, the common ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SECTION: WRITE-UP SHEET 6 OF 41

glass, walls and RCC slabs shall be considered for heat transfer for the duration of time that either of the areas is not air-conditioned. Normally, it is recommended that for adjacent areas, which are separately air-conditioned, the common glass, walls and RCC slabs be considered for heat transfer, for each of the areas. This would give a higher cooling load estimate than that considering the adjacent rooms being air-conditioned. However, the designer has to take an overall system design approach while considering this element of heat gain. (b)

Heat Transfer Coefficient The overall heat transfer coefficient (U-factor) shall be estimated for each of the walls, ceiling and floor slabs and glass areas. The U-factor is the reciprocal of the sum of the total thermal resistance across the slab, wall or glass. The total thermal resistance shall be estimated for each component of the thermal resistance as listed below: (i)

(ii)

Walls •

Outer air film

•

Outer cement plaster

•

Wall (Ordinary brick, hollow concrete block, concrete, plywood etc.)

•

Inner cement plaster

•

Inner air film

Floor Slabs, Roof or Ceiling •

Outer air film

•

Floor finish- tiled finish for floors or water-proofing finish for exposed roof slabs

•

RCC slab

•

Thermal insulation

•

Inner cement plaster

•

Inner air film

Thermal resistances for above shall be referred to in ASHRAE Handbook, Fundamentals 1993 Edition, Chapter 22, Tables 1, 2, 3 ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SECTION: WRITE-UP SHEET 7 OF 41

and 4. The construction materials for the walls, RCC slabs, window or door glass etc. shall be taken from the architectural drawings. If necessary, the civil engineer may be consulted. The above components are listed as guidelines and are generally adequate for most applications. The components for specific applications or projects may be established from the respective architectural drawings. The insulation thickness required to be provided, shall be calculated for prevention of condensation on the external surfaces or the optimised insulation thickness considering the operating and installation costs for the insulation, whichever is more stringent. For exposed roofs, thermal insulation is generally provided out of 50 mm thick expanded polystyrene or equivalent insulation material. In cases where the room temperatures are low enough to cause condensation on the external surface of the room wall, it may be necessary to provide thermal insulation. This requirement of thermal insulation may be verified by the following equation: <

Tr + (Tod – Tr) x (U/U1)

Tod

=

Dew point temperature of the area adjacent to the air-conditioned room

Tr

=

DBT of air-conditioned room

U

=

Overall heat transfer coefficient of the common wall

U1

=

The reciprocal of the common wall thermal resistance considered without the wall outer air film resistance

Tod Where,

The outdoor air dew point temperature shall be taken from the psychrometric chart for the most stringent conditions. Generally, this will be for the monsoon season. In cases where the adjacent area is to be maintained at a high RH, the same shall be considered. Thermal insulation may also be provided for conservation of heat in case of heating applications. This insulation may serve a dual purpose. It may serve as a heat conservator to minimise the heating load during winter and to minimise the heat gain during summer. This will be applicable to sites located in regions having extreme climates. In case of dual duty, the thermal insulation thickness for the more ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

SECTION: WRITE-UP SHEET 8

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

OF 41

stringent requirement shall be provided and considered for the airconditioning load estimation. (iii)

The heat transfer coefficients in Watts /M2 OC for conduction through glass shall be taken from the table 1:

TABLE 1 HEAT TRANSFER COEFFICIENTS FOR CONDUCTION THROUGH GLASS VERTICAL GLASS SINGLE

HORIZONTAL GLASS

DOUBLE

TRIPLE

SINGLE

DOUBLE

SUMMER WINTER SUMMER WINTER AIR SPACE THICKNESS mm

--

6

12.5

20

6

12.5

20

TO

TO

100

100

--

--

6

6

WITHOUT STORM WINDOWS

6.4

3.5

3.1

3.0

2.3

2.0

1.9

4.9

7.9

2.8

4.0

WITH STORM WINDOWS

3.1

3.5

3.1

3.0

2.3

2.0

1.9

2.4

3.6

2.8

4.0

SOURCE: ISHRAE 1997, PART I, TABLE 6, PAGE 1.7

(c)

ETD Table 2 gives the uncorrected ETD (∆tes for wall in shade and ∆tem for wall exposed to the design conditions) values for varying weights of wall. Table 3 gives the uncorrected ETD (∆tes for roof in shade and ∆tem for roof exposed to the design conditions) values for varying weights of the roof structure. Table 4 gives the corrections (∆tc) to be made to the uncorrected ETD (∆tes and ∆tem) from tables 2 and 3 for a variation in the outdoor to indoor DBT difference and daily range. The corrected ETD shall be calculated as per the following equation: ETD

=

∆tesc + {(∆temc - ∆tesc) x c x Rs/Rm}

=

Corrected ETD across the wall or roof (OC)

where, ETD

∆tesc =

Corrected ETD of the roof or the wall under consideration in shade (OC). ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS = ∆temc =

SECTION: WRITE-UP SHEET 9 OF 41

∆tes + ∆tc Corrected ETD of the roof or the wall under consideration subjected to the design condition. i.e. load on the exposed roof or on walls facing the respective directions (0C).

=

∆tem + ∆tc

=

Correction factor for colour shade of wall

=

1 for dark coloured roof or wall

=

0.778 for medium coloured roof or wall

=

0.556 for light coloured roof or wall

Rs

=

Maximum Solar Heat Gain Factor (SHGF) incident on glass or wall or horizontal roof for the month and latitude desired. Rs values shall be taken from ASHRAE Handbook, Fundamentals, 1993 Edition, Chapter 27, Tables 12 to 18. (Watts /M2)

Rm

=

Maximum SHGF incident on glass or wall or horizontal roof for the month of July and latitude of 400 North. Rm values shall be taken from ASHRAE Handbook, Fundamentals, 1993 Edition, Chapter 27, Table 15. (Watts /M2)

c

ISSUE R2 TCE FORM NO. 120R1

2.2 2.2 1.1 1.1 1.1 1.7 2.2 4.4 6.7 8.3 8.9 10.0 10.0 8.3 7.8 6.1 5.6 5.0 4.4 4.4 3.9 3.3 3.3 2.8

-0.6 -1.7 -2.2 -1.7 -1.1 3.9 6.7 11.1 13.3 13.9 14.4 12.8 11.1 8.3 6.7 5.6 4.4 3.3 2.2 1.1 0.6 0.6 0.0 -0.6

293 -0.6 -1.1 -2.2 0.6 2.2 7.8 12.2 15.0 16.7 15.6 14.4 11.1 8.9 6.7 5.6 3.9 3.3 1.7 1.1 0.6 0.6 0.0 0.0 -0.6

98

SOURCE: ISHRAE 1997, PART I, TABLE 9, PAGE 1.14

3.9 3.3 3.3 2.8 2.2 2.2 2.2 2.2 2.2 3.9 5.6 7.2 7.8 8.3 8.9 8.9 7.8 6.7 5.6 5.6 5.0 5.0 4.4 3.9

6 7 8 9 10 11 12 1 2 3 4 5 6 7 8 9 10 11 12 1 2 3 4 5

488 5.0 4.4 4.4 4.4 4.4 3.9 3.3 6.1 7.8 8.3 8.9 10.0 8.9 8.3 7.8 7.2 6.7 6.7 6.7 6.1 6.1 5.6 5.6 5.0

683 3.9 3.9 3.3 3.3 3.3 6.1 8.9 9.4 10.0 10.6 10.0 8.9 7.8 7.2 6.7 6.1 5.6 5.6 5.6 5.0 5.0 4.4 4.4 3.9

488 0.6 0.6 0.0 7.2 11.1 13.3 15.6 14.4 13.9 11.7 10.0 8.3 7.8 7.2 6.7 6.1 5.6 4.4 3.3 2.8 2.2 1.7 1.7 1.1

293

SOUTH EAST

5.6 3.3 7.2 10.6 14.4 15.0 15.6 14.4 13.3 10.6 8.9 8.3 7.8 6.7 5.6 4.4 3.3 2.2 1.1 0.0 -0.6 -0.6 -1.1 -1.1

98 6.1 5.6 5.6 5.0 4.4 5.0 5.6 8.3 10.0 10.6 10.0 9.4 8.9 7.8 6.7 7.2 7.8 7.8 7.8 7.2 7.2 6.7 6.7 6.7

683 2.8 2.8 3.3 4.4 7.8 11.1 13.3 13.9 13.3 11.1 10.0 8.9 7.8 7.8 7.8 7.2 6.7 6.1 5.6 5.0 4.4 3.9 3.9 3.3

488 -0.6 -0.6 0.0 11.7 16.7 17.2 17.2 10.6 7.8 7.2 6.7 7.2 7.8 7.2 6.7 6.1 5.6 4.4 2.8 2.2 1.7 0.6 0.6 0.0

293

EAST

0.6 9.4 16.7 18.3 20.0 19.4 17.8 11.1 6.7 7.2 7.8 7.8 7.8 6.7 5.6 4.4 3.3 2.2 1.1 0.0 -0.6 -1.1 -1.7 -1.7

98 2.8 2.8 3.3 3.3 3.3 3.3 3.3 5.6 7.8 8.9 7.8 6.7 5.6 5.6 5.6 5.6 5.6 5.6 5.6 5.0 5.0 4.4 3.9 3.9

683 2.2 1.7 2.2 2.2 2.2 5.6 8.9 8.3 7.8 6.7 5.6 6.1 6.7 6.7 6.7 6.1 5.6 5.0 4.4 3.9 3.3 3.3 2.8 2.8

488

-0.6 -1.1 -1.1 2.8 13.3 12.2 11.1 8.3 5.6 6.1 6.7 7.2 7.8 7.2 6.7 6.1 5.6 4.4 3.3 2.2 1.1 0.6 0.0 -0.6

293

NORTH EAST

0.0 -1.1 -1.7 -2.2 -1.1

2.8 8.3 12.2 12.8 13.3 10.6 7.8 7.2 6.7 7.2 7.8 7.8 7.8 6.7 5.6 4.4 3.3 2.2 1.1

98

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

AM

PM

AM

683

WT.OF WALL Kg/M 2

SOUTH

TCE.M6-ME-811-301

TIME

EXPOSURE

FOR DARK COLOURED, SUNLIT AND SHADED WALLS; 35°C OUTDOOR DBT; 27°C INDOOR DBT; 11°C DAILY RANGE; 24 HOUR OPERATION; JULY AND 40°N LATITUDE

EQUIVALENT TEMPERATURE DIFFERENCE ∆ tes AND ∆ tem ° C

TABLE 2

TCE CONSULTING ENGINEERS LIMITED SECTION: WRITE-UP

10 OF 41 SHEET

ISSUE R2

TCE FORM NO. 120R1

0.6 0.6 0.0 0.0 0.0 0.0 0.0 0.6 1.1 1.7 2.2 2.8 2.8 2.8 4.4 3.9 3.3 2.8 2.2 1.7 1.7 1.1 1.1 0.6

-1.7 -1.7 -2.2 -1.7 -1.1 -0.6 0.0 1.7 3.3 4.4 5.6 6.1 6.7 6.7 6.7 5.6 4.4 3.3 2.2 1.1 0.6 0.0 -0.6 -1.1

293 -1.7 -1.7 -2.2 -1.7 -1.1 0.6 2.2 4.4 5.6 6.7 7.8 7.2 6.7 5.6 4.4 3.3 2.2 1.1 0.0 0.0 -0.6 -0.6 -1.1 -1.1

98

SOURCE: ISHRAE 1997, PART I, TABLE 9, PAGE 1.14

0.6 0.6 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.6 1.1 1.7 2.2 2.8 3.3 3.9 4.4 3.9 3.3 2.2 1.7 1.1 1.1 0.6

6 7 8 9 10 11 12 1 2 3 4 5 6 7 8 9 10 11 12 1 2 3 4 5

488 4.4 3.9 3.3 3.3 3.3 3.3 3.3 3.3 3.3 3.3 3.3 3.9 4.4 5.0 5.6 7.8 10.0 10.6 11.1 8.9 7.2 6.1 5.6 5.0

683 2.8 2.2 2.2 2.2 2.2 2.2 2.2 2.2 2.2 2.8 3.3 5.0 6.7 9.4 11.1 11.7 12.2 7.8 4.4 3.9 3.9 3.3 3.3 2.8

488 -1.1 -1.7 -2.2 -1.7 -1.1 0.0 1.1 3.3 4.4 5.6 6.7 11.7 16.7 17.2 17.8 11.7 6.7 4.4 3.3 2.2 1.7 0.6 0.0 -0.6

293

NORTH WEST

-1.7 -2.2 -2.2 -1.1 0.0 1.7 3.3 5.6 6.7 10.6 13.3 18.3 22.2 20.6 18.9 10.0 3.3 2.2 1.1 0.0 -0.6 -0.6 -1.1 -1.1

98 6.7 6.1 5.6 5.0 4.4 4.4 4.4 5.0 5.6 5.6 5.6 6.1 6.7 7.8 8.9 11.7 12.2 12.8 12.2 11.1 10.0 8.9 8.3 7.2

683 3.9 3.9 3.3 3.3 3.3 3.3 3.3 3.9 4.4 5.6 6.7 9.4 11.1 13.9 15.6 15.0 14.4 10.6 7.8 6.7 6.1 5.6 5.0 4.4

488 1.1 0.6 0.0 0.0 0.0 1.1 2.2 3.9 5.6 10.6 14.4 18.9 22.2 22.8 20.0 15.6 8.9 5.6 3.3 2.8 2.2 1.7 1.7 1.1

293

WEST

-1.1 -1.7 -2.2 -1.1 0.0 1.7 3.3 7.8 11.1 17.8 22.2 25.0 26.7 18.9 12.2 7.8 4.4 2.8 1.1 0.6 0.0 0.0 -0.6 -0.6

98 4.4 4.4 4.4 4.4 4.4 3.9 3.3 3.3 3.3 3.9 4.4 5.0 5.6 8.3 10.0 10.6 11.1 7.2 4.4 4.4 4.4 4.4 4.4 4.4

683 3.9 2.8 3.3 2.8 2.2 2.8 3.3 3.9 4.4 6.7 7.8 10.6 12.2 12.8 13.3 12.8 12.2 8.3 5.6 5.6 5.0 5.0 4.4 3.9

488

1.1 0.6 0.0 0.0 0.0 0.6 1.1 4.4 6.7 13.3 17.8 19.4 20.0 19.4 18.9 11.1 5.6 3.9 3.3 2.8 2.2 2.2 1.7 1.7

293

SOUTH WEST

0.6 0.6 0.0 -0.6 -0.6

-1.1 -2.2 -2.2 -1.1 0.0 2.2 3.3 10.6 14.4 18.9 22.2 22.8 23.3 16.7 13.3 6.7 3.3 2.2 1.1

98

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

AM

PM

AM

683

WT.OF WALL Kg/M 2

NORTH OR SHADED WALL

TCE.M6-ME-811-301

TIME

EXPOSURE

FOR DARK COLOURED, SUNLIT AND SHADED WALLS; 35°C OUTDOOR DBT; 27°C INDOOR DBT; 11°C DAILY RANGE; 24 HOUR OPERATION; JULY AND 40°N LATITUDE

EQUIVALENT TEMPERATURE DIFFERENCE ∆ tes AND ∆ tem ° C

TABLE 2 (CONTD.)

TCE CONSULTING ENGINEERS LIMITED SECTION: WRITE-UP

11 OF 41 SHEET

ISSUE R2

TCE FORM NO. 120R1

5.6 5.6 5.0 4.4 3.3 2.2 1.1 0.6 0.0 -0.6 -1.1

7 8 9 10 11

12 1 2 3 4

5

-2.8

1.1 0.0 -0.6 -1.7 -2.2

6.1 5.6 4.4 3.3 2.2

4.4 5.6 6.7 7.2 6.7

-2.8

0.0 -0.6 -1.7 -2.2 -2.8

5.6 4.4 2.8 1.1 0.6

6.7 7.2 7.8 7.2 6.7

-1.1 0.0 1.1 3.3 5.0

-2.8 -2.8 -2.2

98

-0.6

3.3 2.2 1.1 0.6 0.0

7.2 6.7 6.1 5.6 4.4

4.4 5.6 6.7 7.2 7.8

-1.1 -1.1 0.0 1.1 2.8

-0.6 -1.1 -1.1

293

-0.6

1.7 0.6 0.0 0.0 -0.6

7.2 6.7 5.0 3.9 2.8

7.2 7.8 7.8 7.8 7.8

-0.6 0.0 1.1 2.8 5.0

-1.1 -1.1 -0.6

195

SOURCE: ISHRAE 1997, PART I, TABLE 10, PAGE

2.2 3.3 4.4 5.0 5.6

2 3 4 5 6

-1.7 -1.1 0.0 1.1 2.8

-2.8 -2.8 -2.2

195

SPRAYED

-1.7

0.0 -0.6 -1.1 -1.1 -1.7

6.7 5.6 3.3 1.1 0.6

10.0 9.4 8.9 8.3 7.8

1.1 2.2 4.4 6.7 8.3

-2.2 -1.1 0.0

98

0.0

3.3 2.2 1.7 1.1 0.6

8.3 7.8 6.7 5.6 4.4

5.6 6.7 7.8 8.3 8.9

-1.1 -1.1 1.1 2.8 3.9

-0.6 -1.1 -1.1

293

-1.7

1.7 0.6 -0.6 -1.1 -1.7

7.8 6.7 5.6 3.9 2.8

8.3 8.3 8.9 8.3 8.3

-0.6 0.0 2.8 5.6 7.2

-1.7 -1.1 -0.6

195

-2.8

0.6 -0.6 -1.1 -1.7 -2.2

6.7 5.6 3.3 1.1 0.6

12.2 11.1 10.0 8.9 7.8

1.1 2.2 5.6 8.9 10.6

-2.8 -1.1 0.0

98

COVERED WITH WATER

7.8

16.7 15.0 12.8 11.1 10.0

20.6 19.4 18.9 18.9 17.8

14.4 15.6 17.8 19.4 20.6

6.1 6.7 7.2 8.9 12.2

7.2 6.7 6.1

391

6.1

13.9 12.2 10.0 8.9 7.2

21.1 20.0 18.9 17.2 15.6

15.0 17.2 19.4 21.1 21.7

3.9 4.4 6.1 8.9 12.2

5.0 4.4 3.3

293

3.3

11.1 9.4 7.2 6.1 5.0

21.7 19.4 17.8 15.6 13.3

15.6 18.3 21.1 22.2 22.8

1.7 3.3 5.6 8.9 12.8

2.2 1.7 1.1

195

1.1

8.3 6.7 4.4 3.3 2.2

22.2 19.4 16.7 13.9 11.1

16.7 20.0 22.8 23.9 23.9

-0.6 1.1 5.0 8.9 12.8

0.0 -0.6 -1.1

98

EXPOSED TO SUN

-1.7

5.6 3.9 1.7 0.6 -0.6

22.8 19.4 15.6 12.2 8.9

17.8 21.1 23.9 25.6 25.0

-2.8 -0.6 3.9 8.3 13.3

-2.2 -3.3 -3.9

49

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

AM

PM

-1.1 -1.1 -0.6 0.0 1.1

-1.7 -1.7 -1.1

6 7 8

9 10 11 12 1

293

WT. OF ROOF Kg/M2

SHADED

TCE.M6-ME-811-301

AM

TIME

CONDITION

FOR DARK COLOURED, SUNLIT AND SHADED ROOFS; 35 °C OUTDOOR DBT; 27°C INDOOR DBT; 11°C DAILY RANGE; 24 HOUR OPERATION; JULY AND 40°N LATITUDE

EQUIVALENT TEMPERATURE DIFFERENCE ∆ tes AND ∆ tem ° C

TABLE 3

TCE CONSULTING ENGINEERS LIMITED SECTION: WRITE-UP

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ISSUE R2

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DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SHEET 13 OF 41

TABLE 4 CORRECTIONS TO ETD ∆ tc ° C OUTDOOR TO INDOOR DBT DIFFERENCE °C

-16.7 -11.1 -5.6 0.0 2.8 5.6 8.3 11.1 13.9 16.7 19.4 22.2

DAILY RANGE °C

4.4

5.6

6.7

7.8

8.9

10.0

11.1

12.2

13.3

14.4

15.6

16.7

17.8

18.9

20.0

21.1

22.2

-21.7 -22.2 -22.8 -23.3 -23.9 -24.4 -25.0 -25.6 -26.1 -26.7 -27.2 -27.8 -16.1 -16.7 -17.2 -17.8 -18.3 -18.9 -19.4 -20.0 -20.6 -21.1 -21.7 -22.2 -10.6 -11.1 -11.7 -12.2 -12.8 -13.3 -13.9 -14.4 -15.0 -15.6 -16.1 -16.7 -5.0 -5.6 -6.1 -6.7 -7.2 -7.8 -8.3 -8.9 -9.4 -10.0 -10.6 -11.1

-28.3 -22.8 -17.2 -11.7

-28.9 -23.3 -17.8 -12.2

-29.4 -23.9 -18.3 -12.8

-30.0 -24.4 -18.9 -13.3

-30.6 -25.0 -19.4 -13.9

-2.2 0.6 3.3 6.1

-2.8 0.0 2.8 5.6

-3.3 -0.6 2.2 5.0

-3.9 -1.1 1.7 4.4

-4.4 -1.7 1.1 3.9

-5.0 -2.2 0.6 3.3

-5.6 -2.8 0.0 2.8

-6.1 -3.3 -0.6 2.2

-6.7 -3.9 -1.1 1.7

-7.2 -4.4 -1.7 1.1

-7.8 -5.0 -2.2 0.6

-8.3 -5.6 -2.8 0.0

-8.9 -6.1 -3.3 -0.6

-9.4 -10.0 -10.6 -11.1 -6.7 -7.2 -7.8 -8.3 -3.9 -4.4 -5.0 -5.6 -1.1 -1.7 -2.2 -2.8

8.9 11.7 14.4 17.2

8.3 11.1 13.9 16.7

7.8 10.6 13.3 16.1

7.2 10.0 12.8 15.6

6.7 9.4 12.2 15.0

6.1 8.9 11.7 14.4

5.6 8.3 11.1 13.9

5.0 7.8 10.6 13.3

4.4 7.2 10.0 12.8

3.9 6.7 9.4 12.2

3.3 6.1 8.9 11.7

2.8 5.6 8.3 11.1

2.2 5.0 7.8 10.6

1.7 4.4 7.2 10.0

1.1 3.9 6.7 9.4

0.6 3.3 6.1 8.9

0.0 2.8 5.6 8.3

SOURCE: ISHRAE 1997, PART I, TABLE 11, PAGE 1.15

The temperature difference for internal surfaces like walls, ceiling, roof or floor slab or glass, shall be the temperature difference between the area to be air-conditioned and the respective adjoining areas. In case the temperature in the adjoining area is not available from the input data, the room shall be assumed to be non-air-conditioned. The temperature difference (∆T) for such cases shall be taken as: ∆T

=

DBT(ambient) - DBT(air-conditioned area)

- 2.8OC

For external glass, the temperature difference shall be taken as: ∆T

=

DBT(ambient) - DBT(air-conditioned area)

For heating coil calculations, the ambient DBT will be lower than the DBT for the air-conditioned area. The temperature difference for all glass area, exposed and internal walls, for heating load calculations shall be the difference between room and ambient DBT. ∆T 4.1.2

=

DBT(air-conditioned area) - DBT(ambient)

Latent Heat Gain This component of heat gain is to be calculated for room design RH values of 45% or lower as the contribution for higher indoor RH values is not significant. The latent heat gain due to moisture ingress shall be estimated by the following equations: ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

(a)

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SECTION: WRITE-UP SHEET 14 OF 41

Moisture Load Through the Fabric of the Structure HLS

=

2.225 x ∆H x V x ∆G x F1 x F2 x F3 x F4

Where,

(b)

HLS

=

Latent load due to vapour permeance through structure (watts)

∆H

=

Enthalpy differential between humidity ratios Gi for room and Go for the adjacent area, at the room DBT from psychrometric chart at altitude of the site (KJ/Kg).

V

=

Room volume (M3)

∆G

=

Differential humidity ratio across the room i.e. difference between outside humidity ratio Go and room humidity ratio Gi (gms of moisture per Kg of dry air)

F1

=

Factor for differential humidity ratio (Refer table 5)

F2

=

Factor for space permeance (Refer table 6)

F3

=

Factor for wall or slab construction (Refer table 7)

F4

=

Factor for vapour treatment (vapour barrier) of walls (Refer table 8)

Moisture Load Due to Door Openings HLD

=

1.357 x ∆H x N x A x ∆G x F1

Where, HLD =

Latent load due to vapour permeance through door openings (watts)

∆H

=

Enthalpy differential between humidity ratios Gi for room and Go for the adjacent area, at the room DBT from psychrometric chart at altitude of the site KJ/Kg.

N

=

Number of door openings per hour

A

=

Area of door opening (M2)

∆G

=

Differential humidity ratio across the door (gms of moisture per Kg of dry air)

F1

=

Factor for differential humidity ratio (Refer table 5) ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

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SECTION: WRITE-UP SHEET

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Moisture Load Through Fixed Openings (Conveyors, Open Windows etc.) HLO

62.039 x ∆H x (A x ∆G x F1) / D

=

Where, HLO

=

Latent load due to vapour permeance through fixed openings (Watts)

∆H

=

Enthalpy differential between humidity ratios Gi for room and Go for the adjacent area, at the room DBT from psychrometric chart at altitude of the site KJ/Kg

A

=

Area of fixed opening (M2)

∆G

=

Differential humidity ratio across the fixed opening (gms of moisture per Kg of dry air)

F1 D (d)

=

Factor for differential humidity ratio (Refer table 5)

=

Depth of the opening (M)

Total Latent Heat Gain HL

=

HLS

+

HLD

+

HLO

TABLE 5 FACTOR F1 FOR DIFFERENTIAL HUMIDITY RATIO DIFFERENTIAL HUMIDITY RATIO (gms per Kg of dry air)

5.00

5.71

7.14

8.57

10.00

FACTOR F1

1.00

1.11

1.35

1.58

1.82

DIFFERENTIAL HUMIDITY RATIO (gms per Kg of dry air)

11.43

12.86

14.29

15.71

17.14

FACTOR F1

2.05

2.29

2.59

2.76

2.99

SOURCE: ENGINEERING DATA MANUAL, BRY-AIR, DEHUMIDIFIERS, TABLE II, PAGE 8.

TABLE 6 FACTOR F2 FOR SPACE PERMEATION VOLUME OF ROOM 0.28 0.57 (IN 1000 M3) FACTOR F2

0.64 0.57

0.85 1.13 1.42 1.70 1.98

2.27 2.55 2.83

0.53 0.50 0.47 0.44 0.42

0.40 0.38 0.37 ISSUE R2 TCE FORM NO. 120R1

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TABLE 6 (CONTD.) VOLUME OF ROOM 3.40 3.96 (IN 1000 M3) FACTOR F2

0.36 0.34

4.53 5.10 5.66 6.23 6.80

7.36 7.93 8.50

0.33 0.32 0.29 0.28 0.27

0.27 0.26 0.25

SOURCE: ENGINEERING DATA MANUAL, BRY-AIR, DEHUMIDIFIERS,TABLE III, PAGE 8.

TABLE 7 FACTOR F3 FOR WALL OR SLAB CONSTRUCTION DETAILS OF CONSTRUCTION

FACTOR F3

Masonry or frame construction Sheet metal, steel welded Module panel, caulked and sealed

1.0 0.2 0.5

SOURCE: ENGINEERING DATA MANUAL, BRY-AIR, DEHUMIDIFIERS, TABLE IV, PAGE 8.

TABLE 8 FACTOR F4 FOR VAPOUR BARRIER DETAILS OF VAPOUR BARRIER Laminated mylar – metallic or polyethylene film Two layers edge sealed moisture paper Two coats vapour proof paint

FACTOR F4 0.50 0.67 0.75

SOURCE: ENGINEERING DATA MANUAL, BRY-AIR, DEHUMIDIFIERS, TABLE IV, PAGE 8.

NOTES:

4.2

1.

If product of F3 and F4 is less than 0.5, use 0.5.

2.

If room is completely vapour proofed with continuous vapour barrier under the floor (or of all-metal welded construction) the factor (F3 x F4) may be reduced to 0.2

FENESTRATION HEAT GAIN (a)

Glass Transmittance Factor (G) This factor considers the type of glass whether single, double etc., the type of shading for the glass whether provided with shading, tints, curtains or blinds etc. This factor shall be referred to in tables 9, 10 and 11. ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

SECTION: WRITE-UP

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

(b)

SHEET 17 OF 41

Glass Solar Heat Gain Factor (SHGF) SHGF shall be selected from ASHRAE Handbook, Fundamentals 1993 edition, Chapter 27, Tables 12 to 18. This component takes into consideration the orientation of the glass with respect to the true north, latitude, time of the day and month of the year. The fenestration heat gain shall be calculated as follows: Qf

=

G x SHGF x A

Qf

=

Fenestration heat gain for the glass (Watts)

G

=

Glass transmittance factor (dimensionless)

Where,

SHGF =

Solar Heat Gain Factor (watts / M2)

A

Area of the glass (M2)

=

The glass transmittance factor shall be referred from tables 9, 10 and 11. TABLE 9 GLASS TRANSMITTANCE FACTOR (G) FOR SOLAR HEAT GAIN THROUGH GLASS WITH AND WITHOUT SHADING TYPE OF GLASS

GLASS FACTOR NO SHADE

INSIDE VENETIAN BLIND 45 o HORIZ. OR VERTICAL OR ROLLER SHADE LIGHT MEDIU COLOUR M COLOUR

DARK COLOUR

OUTSIDE VENETIAN BLIND 45 o HORIZ. SLATS LIGHT LIGHT ON COLOUR OUTSIDE DARK ON INSIDE

OUTSIDE SHADING SCREEN 17 o HORIZ. SLATS MEDIU M COLOUR

OUTSIDE AWNING VENT. SIDES & TOP

DARK LIGHT MEDIU COLOUR COLOUR M OR DARK

ORDINARY GLASS

1.00

0.56

0.65

0.75

0.15

0.13

0.22

0.15

0.20

0.25

REGULAR PLATE 6 mm

0.94

0.56

0.65

0.74

0.14

0.12

0.21

0.14

0.19

0.24

HEAT ABSORBING GLASS 40 TO 48%

0.80

0.56

0.62

0.72

0.12

0.11

0.18

0.12

0.16

0.20

48 TO 56%

0.73

0.53

0.59

0.62

0.11

0.10

0.16

0.11

0.15

0.18

56 TO 70%

0.62

0.51

0.54

0.56

0.10

0.10

0.14

0.10

0.12

0.16 ISSUE R2

TCE FORM NO. 120R1

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18 OF 41

TABLE 9 (CONTD.) DOUBLE PANE ORDINARY GLASS

0.90

0.54

0.61

0.67

0.14

0.12

0.20

0.14

0.18

0.22

REGULAR PLATE

0.80

0.52

0.59

0.65

0.12

0.11

0.18

0.12

0.16

0.20

48 TO 56% ABSORBING OUTSIDE, ORDINARY GLASS INSIDE

0.52

0.36

0.39

0.43

0.10

0.10

0.11

0.10

0.10

0.13

48 TO 56% ABSORBING OUTSIDE, REGULAR PLATE INSIDE

0.50

0.36

0.39

0.43

0.10

0.10

0.11

0.10

0.10

0.12

ORDINARY GLASS

0.83

0.48

0.56

0.64

0.12

0.11

0.18

0.12

0.16

0.20

REGULAR PLATE

0.69

0.47

0.52

0.57

0.10

0.10

0.15

0.10

0.14

0.17

TRIPLE PANE

SOURCE: ISHRAE 1997, PART I, TABLE 5, PAGE 1.7

TABLE 10 OVERALL FACTOR (G) FOR SOLAR HEAT GAIN THROUGH PAINTED GLASS PAINTED GLASS

LIGHT COLOUR

MEDIUM COLOUR

DARK COLOUR

GLASS FACTOR NO SHADE

0.28

0.39

0.50

SOURCE: ISHRAE 1997, PART I, TABLE 5, PAGE 1.7

TABLE 11 OVERALL FACTOR (G) FOR SOLAR HEAT GAIN THROUGH STAINED GLASS STAINED GLASS

AMBER COLOUR

DARK RED

DARK BLUE

DARK GREEN

GREYED GREEN

LIGHT OPALESCEN T

DARK OPALESCEN T

GLASS FACTOR NO SHADE

0.70

0.56

0.60

0.32

0.46

0.43

0.37

SOURCE: ISHRAE 1997, PART I, TABLE 5, PAGE 1.7

4.3

INFILTRATION HEAT GAIN ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SECTION: WRITE-UP SHEET 19 OF 41

The infiltration air flow rate is approximately proportional to the square root of the pressure differential across the gaps. The air conditions vary as the air moves through the various areas or stages of the air-conditioning system. To enable the estimation of the cooling or heating load the air properties are taken at standard air conditions, i.e. at sea level with air pressure of 101.323 KPa and temperature of 21.11 0C. The RH of air is taken as 50%. The major parameters of the air used for the calculation with the values at the standard air conditions are as follows: Specific heat of air

=

1000 J/Kg0C

Density of air

=

1.187 Kg/M3

Latent heat of moisture =

2502.5 KJ/Kg

The cooling load may be calculated from the following formula: =

Cd x A x √( 2∆p/ ç)

F

=

Air flow rate (M3 /sec)

A

=

Area of the gap between door or window and its frame or the crack in the structure (M2)

∆p

=

Pressure drop across the gap (N / M2 )

ç

=

Density of air (Kg/M3)

Cd

=

Discharge coefficient (normally taken as 1.0)

F Where,

The infiltration air adds latent and sensible load to the total cooling load. Generally, the room pressures are higher than the adjacent areas and will indicate a leakage of conditioned air from the room to the outside. Only the air ingress is to be considered for the cooling load estimation. 4.3.1

The sensible heat gain from the infiltration air ingress into the room shall be calculated from the following formula: QFs

=

F x Cp x ç x (To – Ti)

QFs

=

Sensible heat gain due to the infiltrated air ingress (Watts)

F

=

Infiltrated air flow rate (M3 / sec)

Cp

=

Specific heat of air (Joules/Kg0C)

ç

=

Air density (Kg / M3 )

Where,

ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

4.3.2

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

To

=

Outside air DBT (oC)

Ti

=

Inside air DBT (oC)

SECTION: WRITE-UP SHEET 20 OF 41

The latent heat gain from infiltration air ingress into the room shall be calculated from the formula: =

F x ç x ∆H x 1000

QFl

=

Latent heat gain due to the infiltrated air ingress (Watts)

F

=

Infiltrated air flow rate (M3 / sec)

ç

=

Air density (Kg / M3 )

∆H

=

Enthalpy differential between humidity ratios Gi for room and Go for the adjacent area, at the room DBT from psychrometric chart for the altitude of the site (KJ/Kg).

QFl Where,

4.4

INTERNAL HEAT GAIN The internal heat gain includes the following components: (a)

Heat dissipation loads from equipment located within the room. This load includes the sensible and latent loads that may be dissipated by the equipment located within the room. The actual heat dissipated into the room shall be considered for the cooling load estimation. Considering that motor ratings are selected with a margin of about 15 to 20% above the equipment BKW, this load will generally be lower than the connected electrical rating or rated capacity of a driver of a rotating equipment like pump, compressor, spindle or shaft of a machine etc.

(b)

Heat dissipation from the occupancy in the room. This heat dissipation has latent and sensible component. The total heat dissipation is a function of the activity. The proportion of the sensible and the latent components of this total heat dissipation is a function of the temperature in the room. These loads shall be taken from TCE.M1-ME-811-301 - “Basic Study Guide For Air-conditioning Systems”.

(c)

Lighting loads from the lighting system for the room. In some cases special fixtures are provided which also serve as return air inlets. Such fixtures (Troufers) do not reduce the total heat dissipation from the lighting fixture but reduce the lighting heat dissipation into the room. The reduction in the lighting heat dissipation to the room is instead transferred to the return air, reducing the room sensible heat gain and consequently the room dehumidified air flow rate. ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

(d)

4.5

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SECTION: WRITE-UP SHEET 21 OF 41

For heating load calculations the internal heat dissipation loads such as the equipment heat dissipation, lighting loads and occupancy shall not be considered. This will give a higher heating load requirement.

FAN AND DUCT HEAT GAIN The heat gain to the supply air due to supply air fan, duct heat gain shall be estimated as follows: (a)

Fan Heat Gain Qfan

=

Fbkw + B + M

Qfan

=

Fan heat gain (watts)

Fbkw

=

Fan shaft power consumption (watts)

B

=

Fan belt drive losses (normally considered as 3% of fan shaft power consumption) (Watts)

=

0.03 x Fbkw

=

Heat gain due to motor efficiency (η) (Watts)

=

{(1 - η) x Fbkw }/η

where,

M

The above calculation cannot be made for the first estimation of the cooling load. This component therefore shall be taken as a percentage of the estimated total room heat gains (Sum of heat gains as per paras 4.1, 4.2, 4.3 and 4.4) as follows:

(b)

(i)

Upto fan static pressure of 50 mmWG =

7.5%

(ii)

Add 3% for every 25 mmWG increase in fan static pressure

Duct Heat Gain Qduct

=

Uduct x Aduct x ∆T

=

Duct heat gain due to the temperature difference across the duct wall (Watts)

where, Qduct

ISSUE R2 TCE FORM NO. 120R1

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SECTION: WRITE-UP

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SHEET 22 OF 41

Uduct

=

Overall heat transfer coefficient across the duct wall estimated as indicated in para 4.1.1 (b) (watts / M2 0 C)

Aduct

=

Surface area of the duct (M2)

∆T

=

Temperature difference across the duct wall (0C)

The total duct area will not be available for the first estimate of the cooling load calculation. The duct heat gain shall therefore be taken as a percentage of the estimated total room heat gains (Sum of heat gains as per paras 4.1, 4.2, 4.3 and 4.4) as follows: (i) Duct heat gain

=

2.5%

(ii) Air leakage from supply or return air duct

=

2.5%

(iii) Safety and contingencies

=

2.5%

=

15.0%

Hence total fan and duct heat gain (Fan static pressure upto 50 mmWG)

4.6

FRESH AIR HEAT GAIN The fresh air taken into the air-conditioning system adds to the room heat gain due to the “bypass” of the air over the cooling coil of the air handling unit. This “bypass” of the air over the cooling coil is characterised by the “Bypass Factor” (BF). It is a measure of the efficiency of the heat exchange over the cooling coil. The higher the BF the lower is the efficiency of the cooling coil. This factor is dictated by the cooling coil size i.e. the number of rows of the cooling coil and the air velocity across the coil face. The higher the velocity of air across the cooling coil the higher is the BF. On the other hand the higher the number of rows of tubes of the cooling coil across the air flow or the lower the fin spacing, the lower is the BF. The component of the fresh air load to be considered as part of the room load is given by equations listed in paras 4.6.1 and 4.6.2. The balance fresh air load does not constitute a part of the room load. However, this load shall be taken in the total cooling load estimate as indicated in paras 4.6.3 and 4.6.4.

4.6.1

The fresh air sensible heat gain due to the bypass of the fresh air over the cooling coil in the room is given by the following equation: ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

FASH =

SECTION: WRITE-UP SHEET 23 OF 41

F x Cp x ç x (To - Ti) x BF

Where,

4.6.2

FASH =

Fresh Air Sensible Heat (watts)

F

=

Fresh air flow rate (M3 / sec)

Cp

=

Specific heat of air (Joules/Kg 0C)

ç

=

Air density (Kg / M3 )

To

=

Outdoor air DBT (0C)

Ti

=

Room air DBT (0C)

The fresh air latent heat gain due to the bypass of the fresh air over the cooling coil in the room is given by the following equation: FALH =

F x ç x ∆H x BF x 1000

Where,

4.6.3

FALH =

Fresh Air Latent Heat (watts)

F

=

Fresh air flow rate (M3 / sec)

ç

=

Air density (Kg / M3 )

∆H

=

Enthalpy differential between humidity ratios Gi for room and Go for the adjacent area, at the room DBT from psychrometric chart at altitude of the site (KJ/Kg).

The balance fresh air sensible heat gain on the cooling coil is given by the following equation: BFASH

=

F x Cp x ç x (To - Ti) x (1 - BF)

BFASH

=

Fresh Air Sensible Heat (watts)

F

=

Fresh air flow rate (M3 / sec)

Cp

=

Specific heat of air (Joules/Kg 0C)

ç

=

Air density (Kg / M3 )

To

=

Outdoor air DBT (0C)

Where,

ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

Ti 4.6.4

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS =

SECTION: WRITE-UP SHEET 24 OF 41

Room air DBT (0C)

The balance fresh air latent heat gain on the cooling coil is given by the following equation: =

F x ç x ∆H x (1- BF) x 1000

BFALH

=

Fresh Air Latent Heat (watts)

F

=

Infiltrated air flow rate (M3 / sec)

ç

=

Air density (Kg / M3 )

∆H

=

Enthalpy differential between humidity ratios Gi for room and Go for the adjacent area, at the room DBT from psychrometric chart at altitude of the site (KJ/Kg).

BFALH Where,

The values of the BF to be considered for the purpose of cooling load estimation for various applications are as given in table 12. The BF ranges given in table 12 are indicative. For the purpose of the estimation of the cooling load for preparation of specification a conservative value i.e. the highest value given in the range for the particular application may be taken. The correct value of the BF will be available on selection of the cooling coil by the vendor. The fresh air load into the room is a part of the total cooling load due to fresh air, proportional to the BF. The balance of the fresh air cooling load does not constitute a part of the room load. However, this load shall be part of the cooling load imposed on the cooling coil. In case a preliminary cooling coil selection has been done, the values of the BF for finned cooling coil may be taken from table 13. TABLE 12 COOLING COIL BYPASS FACTORS FOR VARIOUS APPLICATIONS COOLING COIL BYPASS FACTOR (BF)

TYPE OF APPLICATION

EXAMPLE

0.30 to 0.50

A small total load or a load that has a low sensible load factor (i.e. high latent load)

Residence

0.20 to 0.30

Typical comfort application with a relatively small total load or a low sensible heat factor with a somewhat larger load

Residence, Small retail shop, Factory

0.10 to 0.20

Typical comfort application

Department store, Bank, ISSUE R2 TCE FORM NO. 120R1

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SECTION: WRITE-UP SHEET 25 OF 41

Factory 0.05 to 0.10

Applications with high internal sensible loads or requiring a large amount of outdoor air for ventilation

Department store, Restaurant, Factory

0.00 to 0.10

100% outdoor air applications

Hospital operation theatres, Clean room, etc.

SOURCE: ISHRAE 1997, PART I, TABLE 14, PAGE 1.16

TABLE 13 COOLING COIL BYPASS FACTORS WITH AND WITHOUT SPRAYS, VARIOUS AIR VELOCITIES AND FIN SPACING DEPTH OF COOLING COIL

WITHOUT SPRAYS 3.1 FINS PER cm

WITH SPRAYS (NOTE 1)

5.5 FINS PER cm

3.1 FINS PER cm

5.5 FINS PER cm

NUMBER OF ROWS AIR VELOCITY M / sec 1.5 to 3.5

1.5 to 3.5

1.5 to 3.5

1.5 to 3.5

BYPASS FACTOR (BF) 2

0.42 to 0.55

0.22 to 0.38

-

-

3

0.27 to 0.40

0.10 to 0.23

-

-

4

0.19 to 0.30

0.05 to 0.14

0.12 to 0.22

0.03 to 0.10

5

0.12 to 0.23

0.02 to 0.09

0.08 to 0.14

0.01 to 0.08

6

0.08 to 0.18

0.01 to 0.06

0.06 to 0.11

0.01 to 0.05

8

0.03 to 0.08

-

0.02 to 0.05

-

SOURCE: ISHRAE 1997, PART I, TABLE 13, PAGE 1.16

NOTES: 1.

Sprayed coils have a lower BF because the spray provides more heat transfer surface for contacting air.

2.

The BFs are to be taken in proportion with the air velocity across the cooling coil.

3.

The number of rows for residences and comfort applications may be taken as 4 rows, whereas those for high latent load applications i.e. for room sensible heat factor values lower than 0.75 may be taken as 6 rows. ISSUE R2 TCE FORM NO. 120R1

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SECTION: WRITE-UP SHEET 26 OF 41

4.

The fin spacing for cooling coils may be taken as 36 fins per cm of cooling coil finned tube length.

5.

The air velocity may be taken as 2.5 M/sec across the cooling coil.

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4.7

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SECTION: WRITE-UP SHEET 27 OF 41

RETURN AIR DUCT HEAT GAIN The return air heat gain shall be estimated using the following equation: =

U x A x ∆T

RAHG

=

Return Air Heat Gain (Watts)

U

=

Overall heat transfer coefficient across the duct wall. U is to be estimated as per para 4.1.1(b) (Watts / M2 0C)

A

=

Area of the duct wall (M2)

∆T

=

Temperature difference across the duct wall (0C)

RAHG where,

4.8

SUMMARY Various heat gains calculated as per paras 4.1 to 4.7 are summarised in the table 14. TABLE 14 SUMMARY OF HEAT GAIN CALCULATION

SL.NO.

ITEM

SENSIBLE HEAT GAIN (Calculated as per para)

LATENT HEAT GAIN (Calculated as per para)

4.1.1

4.1.2

4.2

-

4.3.1

4.3.2

1.

Structural heat gain

2.

Fenestration heat gain

3.

Infiltration heat gain

4.

Internal heat gain

4.4 a, b, c

4.4 a, b

5.

Total room heat

Sl.Nos. 1+2+3+4 = (RSH)

Sl.Nos. 1+2+3+4 = (RLH)

6.

Room Sensible (RSHF)

7.

Fan and duct heat gain

8.

Fresh air bypass heat gain

Heat

TOTAL

RSH RSH+RLH

Factor 4.5 a, b

-

4.6.1

4.6.2 ISSUE R2 TCE FORM NO. 120R1

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SECTION: WRITE-UP SHEET 28 OF 41

TABLE 14 (CONTD.) 9.

Effective room heat

10.

Effective Room Sensible Heat Factor (ESHF)

11.

Balance fresh air heat gain

12.

Return air heat gain

13.

Grand Total Heat (GTH) = GSH + GLH

14.

Grand Sensible Heat Factor (GSHF)

Sl.Nos. 5+7+8 = ERSH

Sl.Nos. 5+7+8 = ERLH ERSH ERSH+ERLH

4.6.3

4.6.4

4.7

-

Sl.Nos. Sl.Nos. 9+11+12 9+11+12 =GSH =GLH GSH GSH+GLH

In cases where the air-conditioning cooling load is to be met by a chilled water or brine system the total refrigeration capacity (Qch) of the chilled water or brine units shall be calculated as follows: Qch

=

GTH x 1.1

The additional 10% load is to account for the pumping and piping heat gain. 5.0

APPARATUS DEW POINT The ESHF is utilised to calculate the Apparatus Dew Point (ADP). This is the temperature that the air would achieve at the outlet of an ideal cooling coil. This temperature is represented by the intersection of the ESHF line and the saturation curve on the psychrometric chart. The ADP is not a temperature that can be physically measured, but is a parameter which is used to estimate the dehumidified air flow rate. The ADP is also used to select the evaporating temperature in case of Direct Expansion (DX) cooling coils or the chilled water supply temperature in case of chilled water cooling coils. The evaporating temperature for DX systems shall be at least 4 to 5 0C lower than the ADP. Whereas, the chilled water cooling coil entering temperature should be at least 3.5 to 4.5 0C lower than the air leaving temperature. Appendix 1 indicates a typical plot of the various sensible heat factors listed in table 14. The ESHF, GSHF, RSHF and ADP are indicated on this representation. In the ISSUE R2 TCE FORM NO. 120R1

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process of calculating the cooling loads the air-conditioning process shall be represented on a psychrometric chart for the altitude of the area being airconditioned. 6.0

DEHUMIDIFIED AIR FLOW RATE The dehumidified air flow rate is given by the equation: DHF

=

ERSH . (1-BF) x (Room DBT – ADP) x Cp x ç

DHF

=

Dehumidified air flow rate (M3/sec)

Cp

=

Specific heat of air (Joules/Kg 0C)

ç

=

Air density (Kg / M3 )

Where,

7.0

REFERENCES

7.1

ASHRAE Handbook, Fundamentals, 1993 Edition

7.2

Indian Society of Heating Refrigerating and Air-Conditioning Engineers – HVAC Handbook, 1997 Edition, Part I – Air-Conditioning

7.3

Engineering Data Manual – BRY-AIR, Dehumidifiers

7.4

Basic Study Guide for Air-conditioning System – TCE.M1-ME-811-301

7.5

Handbook of Air-conditioning System Design – Carrier Air-Conditioning Company, 1965 edition

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SECTION: APPENDIX SHEET 29 OF 41

APPENDIX 1 TYPICAL AIR-CONDITIONING PSYCHROMETRIC PROCESS

ROOM AIR CONDITION APPARATUS DEW POINT (ADP)

DRY BULB TEMPERATURE 0 C

ESHF LINE

RSHF LINE ROOM ENTERING AIR CONDITION

SATURATION CURVE

GSHF

OUTDOOR DESIGN CONDITION

MIXTURE AIR CONDITION ROOM

SPECIFIC HUMIDITY GMS PER KG OF DRY AIR

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DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS APPENDIX 2 SAMPLE AIR-CONDITIONING LOAD CALCULATIONS

1.0

INPUT DATA

SL. NO.

ITEM

PARAMETER

1.

Location of site

New Delhi

2.

Latitude

28.35 0N

3.

Altitude

216 M above MSL

4.

Name of area

Office premises

5.

General arrangement / Arch. drawing

Refer sketch below

6.

Outside design conditions

7.

DBT 0 C

WBT 0 C

Specific Humidity gms/Kg of dry air

6.1

Summer

43.3

23.9

10.70

6.2

Monsoon

35.0

28.3

21.74

6.3

Winter

7.2

5.0

4.60

6.4

Daily range

13.89

Inside design conditions 7.1

Temperature

24 ± 10C DB

7.2

RH

55 + 5 % - No lower limit on RH

7.3

Specific humidity

10.25 gms / Kg of dry air

8.

Lighting loads

15 Watts/M2 of floor area

9.

Equipment loads

Sensible load 1.5 KW, Latent load 0 KW

10.

Occupancy

15 persons - Office activity

11.

Fresh air requirements

1 air change per hour

12.

Exhaust air requirements

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SHEET 31 OF 41

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS APPENDIX 2 (CONTD.)

13.

Filtration

Pre-filters to be provided

14.

Pressurisation

Nil

15.

Air flow pattern

No specific requirements

16.

Hazardous area classification

Safe area

17.

Duration of air-conditioning system

From 0800 Hrs to 1800 Hrs continously

18.

Availability of cooling water / make-up water

No water available

19.

Future expansion plans

No specific plans. Roof is exposed and floor below is not air-conditioned.

PLAN AND SECTION – OFFICE AREA EXPOSED ROOF

3,000

4,500

FALSE CEILING

SECTION AA 6,00

18,000

3,00

12,000

A

A

WINDOW W1 (TYP) (SIZE 4000 X 1200)

GENTS AND LADIES TOILET

OFFICE AREA

3,000

NORTH

3,000

DOOR (TYP)

AIR-CONDITIONING PLANT ROOM

PLAN

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SHEET 32 OF 41

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS APPENDIX 2 (CONTD.)

2.0

AREA CALCULATIONS The area calculations are based on the layout sketch above

SL. NO.

SURFACE

WALL LENGTH M

WALL HEIGHT M

SURFACE FACE AREA M 2

TOTAL

IN ROOM

ABOVE FALSE CEILING

a

UPTO FALSE CEILING

GLASS AREA M 2

WALL AREA M2

b

c

d = a-c

1.

North wall

18.0

4.5

3.0

54.0

27.0

14.4

39.6

2.

West wall

12.0

4.5

3.0

36.0

18.0

4.8

31.2

3.

Part. wall

30.0

4.5

3.0

90.0

45.0

-

90.0

4.

Plant room

5.

6.

4.1

East wall

3.0

4.5

-

13.5

-

-

13.5

4.2

Part. wall

9.0

4.5

-

40.5

-

-

40.5

Exposed roof 5.1

Office area

-

-

-

216.0

-

-

-

5.2

Plant room

-

-

-

9.0

-

-

-

Floor area 6.1

Office area

-

-

-

216.0

-

-

-

6.2

Plant room

-

-

-

9.0

-

-

-

ISSUE R2 TCE FORM NO. 120R1

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SECTION: APPENDIX SHEET 33 OF 41

APPENDIX 2 (CONTD.) 3.0

OVERALL HEAT TRANSFER COEFFICIENT

3.1

External Wall (a)

Outer air film

0.030 Table 1*

(b)

Outer cement plaster 20 mm thick

0.028 Table 4*

(c)

Wall 230 mm thick brick

0.299 Table 4*

(Density 1760 Kg/M3 as per IS 875 Part 1)

3.2

3.3

(d)

Inner cement plaster 15 mm thick

0.021 Table 4*

(e)

Inner air film

0.120 Table 1*

Total thermal resistance (R) of external wall

0.498

Overall heat transfer coefficient (U= 1/R)

2.008 W/M2 °C

Internal Wall (a)

Outer air film

0.120 Table 1*

(b)

Outer cement plaster 15 mm thick

0.021 Table 4*

(c)

Wall 115 mm thick brick (Density 1760 Kg/M3 as per IS 875 Part 1)

0.150 Table 4*

(d)

Inner cement plaster 15 mm thick

0.021 Table 4*

(e)

Inner air film

0.120 Table 1*

(f)

Total thermal resistance (R) of internal wall

0.432

(g)

Overall heat transfer coefficient (U = 1/R)

2.315 W/M2 °C

Exposed Roof (a)

Outer air film

0.030 Table1*

(b)

Water proofing

0.026 Table 4*

(c)

RCC slab 100 mm thick = 0.1 x 0.69

0.069 Table 4* ISSUE R2 TCE FORM NO. 120R1

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SECTION: APPENDIX SHEET 34 OF 41

APPENDIX 2 (CONTD.)

3.4

3.5

(d)

Thermal insulation (t/k) = 0.05/0.034 (Para 6.4 of TCE.M1-ME-811-301 R2)

1.471

(e)

Inner cement plaster 15 mm thick

0.021 Table 4*

(f)

Inner air film

0.160 Table 1*

(g)

Total thermal resistance (R) of roof

1.777

(h)

Overall heat transfer coefficient (U = 1/R)

0.563 W/M2 °C

Floor (a)

Inner air film

0.110 Table 1*

(b)

Floor tile plaster 25 mm thick

0.035 Table 4*

(c)

RCC slab 100 mm thick = 0.1 x 0.69

0.069 Table 4*

(d)

Cement plaster 15 mm thick

0.021 Table 4*

(e)

Tile terrazo 25 mm thick

0.014 Table 4*

(f)

Outer air film

0.110 Table 1*

(g)

Total thermal resistance (R) of floor

0.359

(h)

Overall heat transfer coefficient (U = 1/R)

2.786 W/M2 °C

(a)

Conduction (single glass without storm windows) (Table 1 of this guide)

6.4 W/M2 °C

(b)

Glass transmittance factor (G)

Glass

0.56

(Table 9 of this guide Ordinary glass with light coloured inside venetian blinds) * Table numbers referred are as per ASHRAE Handbook, Fundamentals 1993 Edition, Chapter 22. The outer air film resistance is assumed for wind velocity of 24 KM/Hr and inner film resistance for still air. 3.6

Equivalent Temperature Difference (ETD)

3.6.1

Outdoor and indoor temperature difference (∆T) ISSUE R2 TCE FORM NO. 120R1

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SECTION: APPENDIX SHEET 35 OF 41

APPENDIX 2 (CONTD.) ∆T

=

43.3 – 24

=

19.3°C

For daily range of 13.89°C, the correction to ∆T is interpolated from table 4 of this design guide as follows: ∆T °C

DAILY RANGE °C 13.3

13.89

14.4

7.2

6.93

6.7

16.7 19.3

9.58

19.4

10.0

Correction to ∆T

=

9.68

9.58

Say 9.6°C.

3.6.2

Factor c for colour shade assumed to be for dark walls =

3.6.3

Exposed Walls (a)

9.4

1

North Wall ETD

=

∆tesc + {(∆temc - ∆tesc ) x c x Rs/Rm}

∆tes

=

2.2 °C (From table 2 for wall weight of 488 Kg/M2)

∆tem =

2.2°C (From table 2 for wall weight of 488 Kg/M2)

∆tesc =

2.2 + 9.6°C

=

11.8°C

∆temc =

2.2 + 9.6°C

=

11.8°C

Rs

=

98 W/M2 From Table 14 of ASHRAE Fundamentals, 1993 Edition, Chapter 27.

Rm

=

90 W/M2 From Table 14 of ASHRAE Fundamentals, 1993 Edition, Chapter 27.

Substituting in equation for ETD above: ETD (b)

=

11.8°C

West Wall ISSUE R2 TCE FORM NO. 120R1

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=

SECTION: APPENDIX SHEET 36 OF 41

∆tesc + {(∆temc - ∆tesc ) x c x Rs/Rm} APPENDIX 2 (CONTD.)

∆tes

=

2.2°C (From table 2 for wall weight of 488 Kg/M2)

∆tem =

6.7°C (From table 2 for wall weight of 488 Kg/M2)

∆tesc =

2.2 + 9.6°C

=

11.8 °C

∆temc =

6.7 + 9.6°C

=

16.3°C

Rs

=

679 W/M2 From Table 14 of ASHRAE Fundamentals, 1993 Edition, Chapter 27.

Rm

=

680 W/M2 From Table 14 of ASHRAE Fundamentals, 1993 Edition, Chapter 27.

Substituting in equation for ETD above: ETD (c)

=

16.29°C

say

16.3°C

East Wall ETD

=

∆tesc + {(∆temc - ∆tesc ) x c x Rs/Rm}

∆tes

=

2.2°C (From table 2 for wall weight of 488 Kg/M2)

∆tem =

10.0°C (From table 2 for wall weight of 488 Kg/M2)

∆tesc =

2.2 + 9.6°C

=

11.8 °C

∆temc =

10.0 + 9.6°C =

19.6°C

Rs

=

82 W/M2 From Table 14 of ASHRAE Fundamentals, 1993 Edition, Chapter 27.

Rm

=

83 W/M2 From Table 14 of ASHRAE Fundamentals, 1993 Edition, Chapter 27.

Substituting in equation for ETD above: ETD (d)

=

19.51°C

say

19.5°C

=

∆tesc + {(∆temc - ∆tesc ) x c x Rs/Rm}

Roof ETD

ISSUE R2 TCE FORM NO. 120R1

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DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS ∆tes

SECTION: APPENDIX SHEET 37 OF 41

=

4.4°C (From table 3 for wall weight of 293 Kg/M2)

∆tem =

19.4°C (From table 3 for wall weight of 293 Kg/M2) APPENDIX 2 (CONTD.)

∆tesc =

4.4 + 9.6°C =

14.0 °C

∆temc =

19.4 + 9.6°C =

29.0°C

Rs

=

457 W/M2 From Table 14 of ASHRAE Fundamentals, 1993 Edition, Chapter 27.

Rm

=

459 W/M2 From Table 14 of ASHRAE Fundamentals, 1993 Edition, Chapter 27.

Substituting in equation for ETD above, ETD (e)

say

28.9°C

=

(43.3-24) - 2.8°C

=

16.5 °C

=

(43.3-24) - 2.8°C

=

16.5 °C

Floor ∆T

3.6.4

28.93°C

Partition Walls ∆T

(f)

=

Solar Heat Gain (a)

Glass transmission factor G

=

0.56

(table 9 of this guide)

(b)

SHGF North glass

=

98 W/M2 *

(c)

SHGF West Glass

=

679 W/M2*

(d)

SHGF East Glass

=

82 W/M2*

(*From Table 14 of ASHRAE Fundamentals, 1993 Edition, Chapter 27.) 3.6.4

Infiltration Heat Gain NIL - as room is to be kept under slight positive pressure. ISSUE R2 TCE FORM NO. 120R1

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DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SECTION: APPENDIX SHEET 38 OF 41

Internal Heat Gain (a)

For office worker at 24°C room DBT for office activity the sensible and latent heat dissipation loads from appendix 2 of TCE.M1-ME-811-301 APPENDIX 2 (CONTD.)

(b) 3.6.7

(i)

Sensible heat load

=

75 Watts/person

(ii)

Latent heat load

=

60 Watts/person

Equipment heat dissipation load

=

1.5 KW

Fresh Air Heat Gain Fresh air flow rate

=

1 air change per hour

=

12M width x 18M length x 3M height

=

648 M3/Hr

=

0.18 M3/sec

For a coil bypass factor of 0.165, from table 13 of this design guide, considering a 3 row deep cooling coil, 5.5 fins per cm and an air face velocity on the cooling coil of 2.5 M/sec, the fresh air bypass sensible and latent loads as per paras 4.6.1 and 4.6.2 of this guide will be as follows: FASH

FALH

=

0.18 x 1000 x 1.187 x (43.3 - 24) x 0.165

=

680.4 Watts

=

0.18 x 1.187 x (10.7 - 10.25) x 1000 x 0.165

=

15.9 Watts

The fresh air balance load on the cooling coil given as per paras 4.6.3 and 4.6.4 of this guide will be as follows: BFASH

BFALH

=

0.18 x 1000 x 1.187 x (43.3 - 24) x (1 - 0.165)

=

3443.2 Watts

=

0.18 x 1.187 x (10.7 - 10.25) x 1000 x (1 - 0.165)

=

80.3 Watts

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SECTION: APPENDIX SHEET 39 OF 41

APPENDIX 2 (CONTD.)

TATA CONSULTING ENGINEERS AIR-CONDITIONING LOAD ESTIMATION BY : DATE : CHD. : DATE:

CLIENT : JOB NO. : OFFICE : DEPT. :

TCE 971 DW ME

PROJECT: DOC. NO. : TCE.971-ME-811-01 SUBJECT : HVAC CALCULATION SH : 1 OF 1 REV. NO. : 0

LOCATION : NEW LATITUDE:28.35 ° ALTITUDE: 216M DELHI N DBT WBT RH SP.HUM. ENTHALPY NAME OF THE AREA : OFFICE ( 0 C) ( 0 C) (%) (gm/kg) (J/KG) OUTSIDE 43.3 23.9 10.70 72 FLOOR : FIRST INSIDE 24.0 55 10.25 50 ROOF : EXPOSED UNEXPOSED DIFF. 19.3 0.45 ROOF INSULATION : 50 mm THK EPS or PUF DAILY RANGE (0 C) DESIGN TIME (AM / 4 PM 13.89 PM) ROOM AREA (M2 ) 216 ROOM VOLUME (M3 ) 648 ASSUMED BYPASS FACTOR (BF) 0.165 SELECTED FRESH AIR (M3 /sec) 0.18 GLASS FENESTRATION GAIN TRANSMISSION GAIN - GLASS AREA (M2 ) SOL HEAT G. FACT WATTS AREA (M2 ) TEMP. U. FACT WATTS N GLASS 14.4 98 0.56 790.3 N GLASS 14.4 19.3 6.4 1778.7 S GLASS S GLASS E GLASS E GLASS W GLASS 4.8 679 0.56 1825.2 W GLASS 4.8 19.3 6.4 592.9 NW GLASS NW GLASS NE GLASS NE GLASS SW GLASS SW GLASS SE GLASS SE GLASS SKYLIGHT SKYLIGHT SOLAR & TRANS. GAIN - WALLS AND TRANSMISSION GAIN ROOF N WALL 39.6 11.8 2.008 938.3 PART. 1 90 16.5 2.315 3437.8 S WALL PART. 2 E WALL PART. 3 W WALL 31.2 16.3 2.008 1021.2 PRT.GLASS NW WALL CEILING NE WALL FLOOR 216 16.5 2.786 9929.3 SW WALL FLOOR SE WALL ROOF 216 28.9 0.563 3514.5 INTERNAL HEAT GAIN INFILTRATION HEAT GAIN OCCUP. -15 75 1125.0 LIGHTING -15 X 216 3240.0 EQUIPT. -1.5 1000 1500.0 DESIGNCONDITIONS

SUMMER / MONSOON / WINTER

SUBTOTAL1

10440.0

SUBTOTAL 1 + SUBTOTAL 2 SAFETY FACTOR

SUBTOTAL2

19253.2

29693.2 %

ROOM SENSIBLE HEAT ( RSH ) SUPPLY DUCT HEAT GAIN IN % SUPPLY DUCT LEAK LOSS 15 % OF ST1 & 2

--

APPARATUS DEW POINT TEMPERATURE (0 C )

ESHF = ERSH/ERTH 34827.6 / 35833.5 INDICATED ADP (0 C) SELECTED ADP 0 C 4454.0 ENTHALPY AT ADP (KJ/Kg)

0.97 14.5 41.0 ISSUE R2 TCE FORM NO. 120R1

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FAN HEAT GAIN IN %

COIL ENTERING AIR TEMPERATURE /ENTHALPY 680.4 COIL LEAVING AIR TEMPERATURE /ENTHALPY 34827.6 DEHUMIDIFIED AIR QUANTITY M3 /sec

FRESH AIR SENSIBLE LOAD EFFECTIVE ROOM SENSIBLE HEAT (ERSH)

SECTION: APPENDIX SHEET 40 OF 41

24.9 / 51.1 16.2 / 42.7 3.7

APPENDIX 2 (CONTD.) LATENT HEAT OCCUPANCY 15 x 60 EQUIPMENT SAFETY FACTOR (%) 10 ROOM LATENT HEAT ( RLH ) SUPPLY DUCT LEAK LOSS IN % FRESH AIR LATENT LOAD EFFECTIVE ROOM LATENT HEAT (ERLH) EFFECTIVE ROOM TOTAL HEAT (ERTH) FRESH AIR HEAT SENSIBLE 3443.2 LATENT 80.3 FRESH AIR HEAT SUBTOTAL 3523.5 SUBTOTAL 3

RETURN DUCT HEAT GAIN IN % PLANT ROOM 900 RETURN DUCT LEAK LOSS IN 2% OF SUBTOTAL 3 ------ GRAND TOTAL HEAT ( GTH )

2635.7 787.1 42779.8

90 AC LOAD IN TR = GTH /3520 12.2 990 CHECK FIGURES : ------ COOLING LOAD / UNIT FLOOR AREA 198.05 (WATTS/M2 ) 15.9 DEHUMIDIFIED AIR QUANTITY / UNIT FLOOR 0.0171 AREA (M3 /sec.M2 ) 1005.9 NOTES : 35833.5 PLANT ROOM GAIN: EWALL = 13.5 x 19.5 x 2.008 = 528.6 PART. = 40.5 x 16.5 x 2.315 = 1547.0 ROOF = 9 x 28.9 x 0.563 = 146.4 3523.5 FLOOR = 9 x 16.5 x 2.786 = 413.7 39357.0 Total load 2635.7 WATTS

SUMMARY The summer load for the room is 12.2 TR with a dehumidified air flow requirement of 3.7 M3/sec and an ADP of 14.5 0C

ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-301

DESIGN GUIDE FOR AIR-CONDITIONING LOAD CALCULATIONS

SECTION: APPENDIX SHEET 41 OF 41

APPENDIX 3 ASHRAE PSYCHROMETRIC CHART NO. 1 FOR NORMAL TEMPERATURES AT SEA LEVEL

ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-304

SECTION: TITLE

GUIDE FOR SELECTION OF COOLING AND HEATING COILS

SHEET i OF iv

GUIDE FOR SELECTION OF COOLING AND HEATING COILS

FILE NAMES: M6ME304R2.DOC AND M6ME304R2.DWG

REV. NO.

R0

R1

R2 ISSUE

INITIALS

SIGN.

INITIALS

SIGN.

INITIALS

PPD. BY

RRC

Sd/-

TPR

Sd/-

TPR

CHD. BY

SCM

Sd/-

PRJ

Sd/-

HRK

APD. BY

SJB

Sd/-

RL

Sd/-

RL

SIGN.

INITIALS

SIGN.

R2

DATE

27.12.1989

27.03.2000

20.03.2003 TCE FORM NO. 020 R2 020R2

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-304

GUIDE FOR SELECTION OF COOLING AND HEATING COILS

SECTION: CONTENTS SHEET ii OF iv

CONTENTS SL.NO. 1.0 2.0 3.0 4.0 5.0 6.0 7.0 APPENDICES 1. 2. FIGURES 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. TABLES 1. 2. 3. 4.

TITLE SCOPE USES OF COILS DESIGN AND CONSTRUCTION FEATURES APPLICATION RANGE INPUT AND OUTPUT DATA FOR COIL SELECTION OR SIZING PROCEDURE FOR SELECTION OR SIZING REFERENCES

SH.NO. 1 1 2 8 8 9 10

EXAMPLES OF COOLING COIL SELECTION PROCEDURE EXAMPLES OF HEATING COIL SELECTION PROCEDURE

11 20

TYPES OF FIN-COIL ARRANGEMENTS WATER CIRCUIT ARRANGEMENTS FOR COOLING COILS CHILLED WATER FLOW ARRANGEMENTS TYPICAL REFRIGERANT DISTRIBUTORS COIL CONTROL ARRANGEMENTS FOR MULTIPLE THERMOSTATIC EXPANSION VALVE FED COILS WATER CIRCUIT ARRANGEMENTS FOR HEATING COILS TYPICAL BYPASS FACTORS FOR COILS CHILLED WATER COIL PERFORMANCE CURVES TYPICAL PERFORMANCE CURVES FOR DIRECT EXPANSION COIL WATE PRESSURE DROP WATER VELOCITY FACTOR F2 FACTOR F3 AIR PRESSURE DROP

28 29

RANGES OF STANDARD RATING CONDITIONS PHYSICAL DATA – AIR HANDLING UNITS AIR SIDE RESISTANCE OF COOLING AND HEATING COILS PHYSICAL PARAMETERS OF COIL

39 40 41 42

30 31 32 33 34 35 36 37 37 38 38

ISSUE R2

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-304

GUIDE FOR SELECTION OF COOLING AND HEATING COILS

SECTION: CONTENTS SHEET iii OF iv

CONTENTS (CONTD.) TABLES (CONTD.)

5. 6. 7. 8. 9. 10. 11.

TITLE

K FACTORS FOR CHILLED WATER COILS

SH.NO.

MULTIPLIERS WS FOR WETTED CHILLED WATER COOLING

43 44

COIL SURFACES COIL FACE AREAS (FA) WATER VELOCITY CONSTANT ‘R’ HOT WATER BTU CONSTANTS 2 ROW HOT WATER COIL CAPACITIES STEAM HEATING COIL CAPACITIES

45 46 47 48 49

ISSUE R2

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GUIDE FOR SELECTION OF COOLING AND HEATING COILS

SECTION: REV.STATUS SHEET iv OF iv

REVISION STATUS

REV.NO.

DATE

DESCRIPTION

R0

27.12.1989

---

R1

27.03.2000

Generally revised. Heating coils added.

R2

20.03.2003

Generally revised.

ISSUE R2

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-304

1.0

GUIDE FOR SELECTION OF COOLING AND HEATING COILS

SECTION: WRITE-UP SHEET

1 OF 49

SCOPE Cooling coils for air-conditioning applications are generally used in Air Handling Units (AHUs), fan coil units, packaged air-conditioners, split air-conditioners and window air-conditioners. Steam or hot water heating coils are generally used in AHUs to heat, pre-heat or re-heat air whereas for smaller capacity heating applications such as in window air-conditioners, electric resistance heating coils are used. This document describes various types, design and construction features of cooling and heating coils for air-conditioning applications and provides guidelines for selection and sizing of these coils.

2.0

USES OF COILS

2.1

COOLING COILS

2.1.1

Cooling coils are generally used for air cooling with or without accompanying dehumidification. Examples of cooling applications without dehumidification are pre-cooling coils, sensible cooling coils or coils that use ordinary or cooling tower water at relatively high temperatures to reduce refrigeration capacity. However, a sizeable portion of coil applications involves simultaneous air sensible cooling and dehumidification.

2.1.2

The usual cooling media used in extended surface coils are chilled water or volatile refrigerants. Brines are seldom used for air-conditioning applications but could be used for some process air cooling or low humidity and low temperature applications or sometimes when brine from industrial system already installed is the only source of refrigeration.

2.1.3

Generally, for air-conditioning applications the temperature of cooling medium such as chilled water or brine or volatile refrigerants like R22, R134a, R123, ammonia etc. are sufficiently above 0 0 C and no frosting is expected in normal operation. For coils operating with cooling medium temperature below 0 0 C, a defrosting mechanism has to be provided. The method of defrosting may be hot gas, electric heaters, water sprays or air.

2.2

HEATING COILS

2.2.1

Heating coils are generally used for temperature and humidity control. Normally, the heating media used in coils is steam or hot water.

2.2.2

Heating coils may be used in AHUs as pre-heating or re-heating coils in conjunction with cooling and dehumidifying coils. These may also be installed in supply air ducts to air-conditioned areas for re-heating. Sometimes these coils are used in fresh air supply system as pre-heating coils.

ISSUE R2 TCE FORM NO. 120R1

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2.2.3

Direct electric resistance air heating coils, also known as blast coils or duct heaters may be used for the same applications as steam or hot water coils. These may be used in smaller sizes and where flow connections or maintenance of pitch in piping would be troublesome. Electric coils are sometimes installed in branch ducts of central fan system and hot water systems to provide the final temperature and relative humidities required for comfort or process air-conditioning. Electric coils installed for use primarily in the heating cycle can also be utilised for re-heat in the cooling cycle if located downstream of the cooling coil. When using electric coils safety aspects like fire hazards should be considered.

3.0

DESIGN AND CONSTRUCTION FEATURES Coils are basically of two types, those consisting of bare tubes and those having extended or finned surfaces. The former is seldom times used where conditions can cause frost accumulation or for cooling within sprayed coil dehumidifiers. The design and arrangement of a coil, constructed with extended surface on the air side, involves consideration of various parameters such as : (a)

Air face velocity

(b)

Material of fins and tubes

(c)

Types and spacing of fins

(d)

Tube diameters and nesting centre dimensions

(e)

Surface ratio

(f)

Provisions to increase air turbulence

(g)

Chilled water flow or refrigerant distribution for cooling coils and hot water flow or steam distribution for heating coils.

3.1

AIR FACE VELOCITY Air flow rate, M3 /Hr

Air face velocity, M/sec = Coil face area, M2 x 3,600 3.1.1

Cooling Coil The extended surface coils may be designed with air face velocity of 1.5 to 4 M/sec. Higher air face velocities give better heat transfer characteristics with higher pressure drop and higher coil bypass. Higher velocities used with dehumidifying coils may also lead to condensed moisture carry-over in ducting system downstream of the coil. As a good compromise between various factors dehumidifying coils are normally designed for an air face velocity of around 2.5 M/sec. However, sensible cooling coils may be designed for air face velocity as high as 4 M/sec. ISSUE R2 TCE FORM NO. 120R1

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-304

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SECTION: WRITE-UP SHEET

3 OF 49

Heating Coil The extended surface coils may be designed with air face velocity of 1 to 4 M/sec. Higher air face velocities give better heat transfer characteristics with higher pressure drop. When heating coils are used in conjunction with cooling and dehumidifying coils in AHUs, the air face velocity is kept same as for cooling and dehumidifying coil. For duct mounted or independent heating coils higher air face velocities may be used.

3.2

MATERIAL OF FINS AND TUBES

3.2.1

Cooling Coil Cooling coils for water or volatile refrigerants usually have aluminium fins and copper tubes. Copper tubes and copper fins are also used quite frequently in corrosive atmosphere like marine applications. Copper tubes and fins also minimise galvanic corrosion and loss of fins bonding to the tube, extending the life of the coil. Aluminium tubes with aluminium fins are also finding usage. Although copper tubes with copper fins are expected to give best heat transfer characteristics most commonly available indigenous designs are with copper tubes and flat or configurated plate aluminium fins. For ammonia refrigeration system normally carbon steel tubes are used since ammonia reacts with copper.

3.2.2

Heating Coil Copper and aluminium are the materials most commonly used in the fabrication of extended surface heating coils. Tubing made of steel or various copper alloys is used in applications where corrosive forces might attack the coils from either inside or outside. The most common combination for low-pressure applications is aluminium fins on copper tubes. Low-pressure steam coils are usually designed to operate upto 10 to 13 Kg/cm2 (g). Above this pressure, tube materials such as red brass, admiralty brass or cupronickel is selected. It is preferable to operate steam heating coil below 3.5 Kg/cm2 (g), as otherwise IBR approval is required.

3.3

TYPES AND SPACING OF FINS

3.3.1

Numerous types of fin arrangements are used such as smooth spiral, flat plate and configurated plate fins as shown in Figure 1. The spiral fins surround each tube individually in all cases, while the plate type may be either continuous including several rows of tubes or individual fin for each tube in round or square shape. Most common designs available in India are flat plate or configurated plate continuous fins.

3.3.2

A major factor in performance of extended surface coils is the bond between the fin and the tube. An intimate contact must be permanently maintained to assure unimpeded heat transfer. ISSUE R2 TCE FORM NO. 120R1

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3.3.3

In individually finned tubes, fins are generally wound on the tube under pressure, to upset the metal slightly at the fin root and are then given a coating of solder while still revolving to assure a uniform coat. In some designs the spiral fins are knurled into a shallow groove on the tube exterior. In very common flat or configurated continuous plate fin designs, the tubes are usually mechanically expanded after the fins are assembled or tube hole collars of a fin are made to override those in the preceding fin thus compressing the fins on the tubes. Most designs available indigenously are of the two types described last.

3.3.4

Various coil designs are made with fins spaced from 3 to 14 per 25 mm. However, most coils used in air-conditioning applications are with 8 to 14 fins per 25 mm. Although the coils with densely packed fins are expected to give better heat transfer, fin spacing should be selected according to job to be performed with special attention given to practical considerations such as condensate drainage, possibility of lint and dust product accumulation and especially at lower temperatures, frost accumulation. Based on above, for normal comfort airconditioning, one may select the coils with 14 fins per 25 mm, while applications such as textile processing industry, pharmaceutical formulation plants (granulating and tabling sections) may require wide spaced fins say 8 fins per 25 mm. Applications involving possibility of frost formation such as cold diffusers may call for still wider fin spacing, say 3 to 5 fins per 25 mm.

3.4

TUBE DIAMETERS AND NESTING CENTRE DIMENSIONS

3.4.1

The cooling and heating coils are built with tubes of commonly 6.4 mm(1/4”), 9.5 mm(3/8”), 12.7 mm(1/2”), 15.9 mm(5/8”), 19.1 mm(3/4”) and 25.4 mm(1”) outside diameter. However, most commonly available designs are with 12.7 and 15.9 mm outside diameter tubes. The tube nesting centres generally vary from 15.9 to 63.5 mm for cooling coils and 25 mm to 75 mm for heating coils. Spacing for 15.9 mm outside diameter tubes is usually 38 to 50 mm vertical and 27 to 38 mm horizontal. Coils with closely spaced tubes are expected to result in better-finned surface effectiveness.

3.4.2

Due to high temperature difference between the heating medium (i.e. hot water or steam) and air, one or two tube rows are normally sufficient for heating applications. For Direct Expansion (DX) coils, a three or four row deep coil is the most common coil used for comfort air-conditioning applications. A five row DX coil may be required for applications having a large percentage of latent load. For chilled water coils, three, four or five rows are used for normal air-conditioning applications. Six row coils are generally used for very large percentage latent load applications.

ISSUE R2 TCE FORM NO. 120R1

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3.5

GUIDE FOR SELECTION OF COOLING AND HEATING COILS

SECTION: WRITE-UP SHEET

5 OF 49

SURFACE RATIO The surface ratio is defined as ratio of total external heat transfer surface of coil to that of internal surface of tubes. Most common designs of extended surface coils usually have surface ratios ranging from 18 to 34.

3.6

PROVISIONS TO INCREASE AIR TURBULENCE Air turbulence and hence outer surface heat transfer could be improved in extended coils by using configurated fins rather than flat fins and staggered tube arrangement as against in-line tube arrangements. However, these features also add to air side friction resistance, though slightly.

3.7

CHILLED AND HOT WATER FLOW & REFRIGERANT DISTRIBUTION

3.7.1

Chilled Water Flow Distribution

AND STEAM

To ensure sufficient water velocities inside tube for adequate heat transfer and at the same time to avoid excessive pumping head, water coils can be provided with various water circuit arrangements. For instance a typical 12 tubes high and six tubes deep coils can be arranged in 3 circuits (quarter-circuited), 6 circuits (halfcircuited), 12 circuits (full-circuited) or 24 circuits (double- circuited) parallel water circuits as shown in Figure 2. The circuits could be arranged to get water velocities usually ranging between 0.3 and 2.4 M/sec and pressure drop across coils varying between 1.5 and 15 MWC. Common designs are usually with 1 to 2 M/sec and pressure drop of around 7 MWC. The normal air-conditioning system designs usually employ full circuited coils while half or quarter circuited coils may find application for sensible cooling coils operating with large Log Mean Temperature Difference (LMTD). Double-circuited coils may have application in all fresh air cooling coils handling relatively smaller volumes of air with large tonnages. The chilled water flow through coil can be in parallel or counter flow arrangement as shown in Figure 3. Counter flow is most preferred arrangement for normal coils to secure advantage of highest possible mean temperature difference between air and chilled water. However, parallel flow arrangement may find usage while sizing coil with large terminal temperature difference between leaving air and leaving chilled water. 3.7.2

Refrigerant Distribution (DX Coils) Coils using volatile refrigerants present more complex cooling fluid distribution problems than water or brine coil. DX coils are used on two types of refrigeration systems i.e. flooded and dry expansion.

ISSUE R2 TCE FORM NO. 120R1

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SECTION: WRITE-UP SHEET

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The flooded system is used mainly for low temperature applications where a small temperature difference between the air and refrigerant is required. However, this system needs large volume of refrigerant. For dry expansion coils two of the most commonly used refrigerant liquid metering devices are the capillary tube and the thermostatic expansion valve. Capillary tube is frequently used with factory built self-contained air-conditioning units upto maximum capacity of 35.2 KW (10 TR) while it is extensively used in India for small capacity models such as window air-conditioners and split units. The thermostatic expansion valve is commonly used for all larger factory assembled units as well as all field assembled central AHUs. Internally equalised thermostatic expansion valves are used for capacity upto about 5 TR. Beyond this capacity, it is recommended that externally equalised thermostatic expansion valves be used. Thermostatic expansion valve maintains a constant superheat for the leaving refrigerant gas. For details of thermostatic expansion valves, refer ASHRAE Handbook, Refrigeration. To ensure reasonably uniform refrigerant distribution in multi-circuit coils, distributing means are provided between expansion valve and the coil to divide refrigerant equally among all coil circuits. Such a refrigerant distributor must be effective in distributing both liquid and vapour, because the entering refrigerant is a mixture of the two. Figure 4 shows five typical types of refrigerant distributors. (a)

In refrigerant distributor 'A', the liquid vapour mixture from the thermostatic expansion valve is led tangentially into a chamber. The coil circuit connections extend outward radially at the top of this chamber.

(b)

In distributor 'B', the refrigerant is discharged at a high velocity through an orifice against the end plate, forming a uniform mixture of vapour and liquid within the refrigerant distributor, from which individual connections are led.

(c)

In distributor 'C' the refrigerant enters at high velocity from the thermostatic expansion valve and is discharged against sidewall to operate a liquid vapour mixture, which is led to the end plug in which the individual circuit connections are closely arranged.

(d)

'D' and 'E' are typical pressure type distributors. The venturi throat in 'D' minimises pressure loss. The liquid vapour mixture flows into the orifice, which is selected to match the load so that adequate mixing of the two phases is maintained to feed equal portions of the mixture to individual circuits. Beyond the orifice, a conical part deflects the refrigerant stream towards the individual circuits. The distributors can be used in either vertical or horizontal position. However, operation in vertical position leads to best distribution.

ISSUE R2 TCE FORM NO. 120R1

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Large coils fed with multiple thermostatic expansion valves can be provided with two arrangements for partial load control i.e. face control and depth control as shown in Figure 5. Face control is used in most cases and provides substantially equal loading on various refrigerant circuits. However, this has the disadvantage of permitting re-evaporation of condensate on the coil and bypassing air into conditioned space during partial load operation. Depth control, seldom available as standard equipment, is usually supplied on special orders. Depth control usually produces unequal loading on the thermostatic expansion valve. Hence, care shall be taken to properly design and size the thermostatic expansion valves to overcome this unbalance. Normally, depth control would be required for deep row coils having six or more rows used for high latent load applications. Depth control eliminates air bypassing during partial load operation and minimises condensate reevaporation. 3.7.3

Hot Water Flow Distribution To obtain proper water velocity in tubes for most efficient heat transfer capacity, without excessive pressure drop through the coil, various circuit arrangements are used as shown in Figure 6. The most common circuiting arrangement is full circuiting, which is considered as standard circuiting. In this arrangement, all tubes in each coil row are supplied with an equal amount of water through a manifold commonly called coil header. A single tube serpentine circuit arrangement can be used on small size booster heater requiring small water quantities upto a maximum of approximately 0.25 to 0.32 lit/sec. With this arrangement a single tube handles the entire water quantity.

3.7.4

Steam Distribution For proper performance of steam heating coils, condensate must be rapidly eliminated and steam must be uniformly distributed to the individual tubes. Condensates are removed by providing proper steam traps and condensate drain lines. Uniform steam distribution is accomplished by different methods such as : (a) Individual orifices in the tubes (b)

Distributing plates in the steam headers and

(c)

Special perforated small diameter inner steam distributing tubes extending into the larger diameter tubes of the primary surface. Coils of the perforated inner tube type are constructed with different arrangements such as : (a) Supply and return on one end, with the incoming steam used to heat the leaving condensate ( non-freeze type) (b)

Supply and return on opposite ends and

(c)

Supply and return on one end and a supply on the opposite end( non-freeze type).

ISSUE R2 TCE FORM NO. 120R1

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Properly designed and selected steam distribution tube coils distribute the steam throughout the entire length of all primary tubes, even when the leaving air temperature is controlled by modulating the steam supply through a steam metering valve. As steam is condensing, the steam temperature is constant. The LMTD is not dependent on direction of steam flow. Steam and air in steam heating coils are in cross flow only. 4.0

APPLICATION RANGE As per ARI 410-2001, cooling coil and heating coil standard rating shall be shown within the range of operating conditions in Table-1. For special applications the range in above design variables could be exceeded.

5.0

INPUT AND OUTPUT DATA FOR COIL SELECTION OR SIZING

5.1

INPUT DATA Following input data is required for selection and sizing of coils :

5.1.1

5.1.2

Cooling Coil (a)

Air flow rate

(b)

Entering air DBT and WBT

(c)

Leaving air DBT and WBT

(d)

Total cooling load

(e)

Sensible cooling load

(f)

Entering chilled water temperature

(g)

Maximum leaving chilled water temperature

(h)

Maximum allowable water side and air side pressure drop

(i)

Assumed bypass factor in arriving at the air flow rate

Heating Coil (a)

Air flow rate

(b)

Entering air DBT

(c)

Leaving air DBT

ISSUE R2 TCE FORM NO. 120R1

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SECTION: WRITE-UP SHEET

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(d)

Total heating load

(e)

Entering hot water pressure, temperature and available water flow rate or desired temperature drop (for hot water coils)

(f)

Steam pressure and temperature (for steam coils)

(g)

Maximum allowable water side and air side pressure drop

OUTPUT DATA The following are the output data from a cooling or heating coil selection : (a)

Coil face area and overall dimensions

(b)

Coil cooling or heating medium flow rate

(c)

Coil cooling or heating medium temperature rise or drop

(d)

Coil cooling or heating medium pressure drop

(e)

Coil air side pressure drop

(f)

Coil air outlet DBT and WBT

(g)

Coil spin spacing, tube size, number of rows.

6.0

PROCEDURE FOR SELECTION OR SIZING

6.1

COOLING COILS

6.1.1

Unit Rating Method Coils can be sized or selected from manufacturers’ data furnished for specific unit size of given face area and rows depth. Typical examples of sizing coils by this method are illustrated for chilled water coil and DX coil in examples 1 and 3 respectively of Appendix-1.

6.1.2

Basic Data Method Certain manufacturers furnish derived heat transfer parameters in the form of tables or charts from which by taking certain physical dimensions of coil face area the coil rows depth is determined. In this procedure it is frequently necessary to re-check and re-select the various parameters to more closely match the required air and cooling fluid conditions with integral rows depth. Typical example of sizing coil by this method is illustrated in example 2 of Appendix-1.

ISSUE R2 TCE FORM NO. 120R1

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SECTION: WRITE-UP SHEET

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HEATING COILS The selection of heating coils is relatively simple, because it involves DBT and sensible heat only. Heating coils are usually selected from manufacturers’ performance charts or tables giving final air temperature at various air face velocities, entering air temperatures and steam or water temperatures. Most coil manufacturers have their own methods of producing performance rating charts or tables from a suitable number of coil performance tests. Typical example of selecting coil from manufacturer’s charts, table or curves is given in Appendix-2.

6.3

COIL SIZING FROM FUNDAMENTAL PRINCIPLES By this method coil can be sized by estimating fundamental heat transfer resistances. This method helps in sizing coils without much relying on manufacturers’ data or establishing the performance of already installed coil with varying parameters. This method also needs few iterations to finalise the exact parameters. This method is normally not used in jobs executed by TCE.

6.4

COMPUTER SELECTION Some manufacturers use computer programmes to select the cooling and heating coils. Input requirements are the same as given in para 5.1. Computer output gives the various parameters of the selected coils.

7.0

REFERENCES

7.1

ASHRAE Handbook, Equipment, 1979 Edition

7.2

ASHRAE Handbook, Refrigeration, 2002 Edition

7.3

ARI 410-2001 Standard for Forced-Circulation Air-Cooling and Air-Heating Coils

7.4

Engineering Bulletin - Hot water heating coils - Blue Star Limited

7.5

Engineering Bulletin - Air Handling Units - Frick India Limited

7.6

Air-Conditioning & Refrigeration Engineer’s Handbook - Voltas Limited, 1977 Edition

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SHEET 11 OF 49

APPENDIX-1 EXAMPLES OF COOLING COIL SELECTION PROCEDURE

The following abbreviations and suffixes are used in Appendix-1 and Appendix-2. Units are indicated at appropriate places. ABBREVIATIONS Q V t h ∆p

= = = = =

Flow rate Velocity Temperature Enthalpy Pressure drop

SUFFIXES 1 = entering 2 = leaving o = outside i = inside a = air

∆h ∆t BF FA GTH

= = = = =

Difference in enthalpy Difference in temperature Bypass Factor Face Area Grand Total Heat

af adb awb cw hw

= = = = =

air face air dry bulb air wet bulb cold water hot water

EXAMPLE-1: CHILLED WATER COIL - BASED ON RATING CHARTS FURNISHED BY VOLTAS LIMITED 1.0

INPUT DATA (a)

Air flow rate, Qa

=

14,000 CFM

(b)

Entering air DBT, tadb1

=

79 0 F

(c)

Entering air WBT, tawb1

=

68 0 F

(d)

Leaving air DBT, tadb2

=

59 0 F

(e)

Leaving air WBT, tawb2

=

58 0 F

(f)

Entering chilled water temperature , t cw1

=

46 0 F

(g)

Allowable chilled water pressure drop, ∆pcw

=

7 MWC

(h)

BF (Assumed during estimation of air flow rate)

=

0.10

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SECTION: APPENDIX SHEET 12 OF 49

APPENDIX-1 (CONTD.) 2.0

SELECT COIL FACE AREA Minimum coil face area to keep the air face velocity around 500 FPM (2.5 M/sec) Coil face area, FA = Qa / 500

= 14,000 / 500 = 28 FT2

From Table-2 select AH-11 having nearest face area to meet the requirement i.e. 34 FT2 (3.16 M2 )

Air face velocity, Vaf = 14,000 / 34 = 411.76 FPM Say 3.0

= 410 FPM (Acceptable)

CHECK BYPASS FACTOR From Figure-7, select 4 rows depth coil as it has bypass factor of about 0.08 at 410 FPM, which is reasonably close to assumed bypass factor of 0.1.

4.0

ESTIMATE CHILLED WATER FLOW RATE Enthalpy of entering air, ha1

= 32.50

Enthalpy of leaving air, ha2

= 25.25 BTU/lb (*) --------------= 7.25 BTU/lb

(*)

ha1-ha2 = ∆h From psychrometric chart

Grand Total Heat (GTH)

BTU/lb (*)

= 4.5 x Qa x ∆h = 4.5 x 14,000 x 7.25 = 4,56,750 BTU/Hr

Say

= 4,57,000 BTU/Hr (tadb1 - tadb2 )

Apparatus Dew Point (ADP) =

tadb1 (1 - BF) ISSUE R2

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TCE.M6-ME-811-304

SECTION: APPENDIX SHEET 13 OF 49

APPENDIX-1 (CONTD.) (79 - 59)

Say

=

79 -

=

(1 - 0.08) 57.26 F

=

57 0 F

0

GTH Coil loading rate (q)

= (ADP – tcw1 ) 4,57,000 = (57-46) =

41,545 BTU/Hr 0 F

As indicated in Figure-8, Chart q = Actual q / Factor.

Factor for Unit AH-11 is 3.

41,545 Chart q

=

=

13,848 BTU/Hr0 F

3 Say

=

13,900 BTU/Hr0 F

From Figure-8, read chilled water flow rate at q = 13,900 BTU/Hr 0 F and air face velocity

= 410 FPM

Chart chilled water flow rate

= 23 USGPM

Actual GPM = Chart GPM x 3.

Factor for Unit AH-11 is 3

Actual flow rate, Qcw =

23 x 3

= 69 USGPM = 15.68 M3 /Hr

Say 5.0

= 16 M3 /Hr

ESTIMATE WATER SIDE PRESSURE DROP From Figure-8 top curve read water side pressure drop for unit flow of 23 USGPM as 2 psi ISSUE R2

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-304

GUIDE FOR SELECTION OF COOLING AND HEATING COILS

SECTION: APPENDIX SHEET 14 OF 49

APPENDIX-1 (CONTD.) Actual psi = Chart psi x Factor.

Factor for Unit AH-11 is 0.92

Pressure drop referred to actual coil size = 2 x 0.92 = 1.84 psi = 1.84 x 0.7 MWC Chilled water pressure drop, ∆pcw 6.0

= 1.29 MWC

Say 1.3 MWC

ESTIMATE AIR SIDE PRESSURE DROP From Table-3 read air side pressure drop for 4 row wet coil corresponding to 410 FPM air face velocity and wet surface coil (interpolated value) as Pressure drop

7.0

=

∆pa

=

0.384 inches of WC

=

9.75 mmWC

ESTIMATE CHILLED WATER TEMPERATURE RISE GTH

Chilled water temperature rise , ∆tcw =

Qcw (GPM) x 500 = 4,57,000 / (69 x 500) = 13.25 0 F = 7.36 0C This temperature is higher than the maximum temperature rise of about 6 0 C generally specified. If the temperature rise is higher than that specified, another coil has to be selected meeting other requirements and temperature rise brought within the specified limit. 8.0

SUMMARY OF PARAMETERS Coil face area ,FA

= 3.16 M2 (34 FT2 )

Total cooling capacity

= 38 TR (4,57,000 BTU/Hr)

Chilled water flow rate, Qcw = 16 M3 /Hr ISSUE R2

TCE CONSULTING ENGINEERS LIMITED GUIDE FOR SELECTION OF COOLING AND HEATING COILS

TCE.M6-ME-811-304

SECTION: APPENDIX SHEET 15 OF 49

APPENDIX-1 (CONTD.) Number of rows deep

= 4

Fin spacing

= 5 fins/cm (13 fins/inch)

Water side pressure drop, ∆pcw = 1.3 MWC Air side pressure drop, ∆pa

= 9.75 mmWC

Based on the layout constraints and meeting the coil face area requirements, manufacturer shall furnish coil dimensions.

EXAMPLE-2: CHILLED WATER COIL - BASED ON BASIC DATA FURNISHED BY FRICK INDIA LIMITED 1.0

INPUT DATA Same as for example 1.

2.0

SELECT COIL FACE AREA Minimum coil face area to keep air face velocity around 500 FPM (2.5 M/sec) Coil face area, FA

= Qa / 500 = 14,000 / 500 = 28 FT2

From Table-4 select model FAH-145 having face area of 29 FT2 and 28 tubes in face. Actual air face velocity, Va

= 14,000/29 = 482.76 FPM

Say 3.0

= 480 FPM

ESTIMATE CHILLED WATER FLOW RATE Assume temperature rise through coil (∆tcw) = 10 0 F GTH Chilled water flow rate, Qcw =

500 x ∆tcw ISSUE R2

TCE CONSULTING ENGINEERS LIMITED GUIDE FOR SELECTION OF COOLING AND HEATING COILS

TCE.M6-ME-811-304

SECTION: APPENDIX SHEET 16 OF 49

APPENDIX-1 (CONTD.) 4,57,000 = 500 x 10 = 91.4 USGPM Say

= 90 USGPM

Actual temperature rise, ∆tcw =

91.4 x 10 90

= 10.16 0 F Leaving chilled water temperature, tcw2 = tcw1 + ∆tcw = 46 + 10.16 = 56.16 0 F

4.0

ESTIMATE WATER VELOCITY THROUGH TUBES Assuming full circuited coil Qcw(USGPM) x 1.235 Water velocity Vcw

=

Number of tubes in face where 1.235 is the factor for 5/8”diameter coil 90 x 1.235 = 28

Say 5.0

=

3.97 FPS(Acceptable as within applicable range)

=

4 FPS

ESTIMATE SENSIBLE HEAT TRANSFER FACTOR (K) MULTIPLYING FACTOR FOR WETTED SURFACE COIL (Ws)

AND

From Table-5 read value of K for water velocity of 4 FPS and air face velocity of 480 FPM (interpolated) as K Coil Sensible Load

=

173.20 BTU/(Hr row FT2 of coil face area 0 F LMTD)

=

1.08 x Qa (tadb1 - tadb2 ) ISSUE R2

TCE CONSULTING ENGINEERS LIMITED GUIDE FOR SELECTION OF COOLING AND HEATING COILS

TCE.M6-ME-811-304

SECTION: APPENDIX SHEET 17 OF 49

APPENDIX-1 (CONTD.) = 1.08 x 14,000 x (79-59) = 3,02,400 BTU/Hr Coil Sensible Load Sensible Heat Factor (SHF) =

3,02,400 =

GTH

4,57,000

= 0.6617 Say = 0.66 From Table-6 read value of Ws for SHF = 0.66 and water velocity = 4 FPS Ws

= 0.891 (tadb1 - tcw2 ) - (tadb2 - tcw1 )

LMTD

= (tadb1 - tcw2 ) ln (tadb2 - tcw1 ) (79 - 56.16) - (59 - 46) = (79 - 56.16) ln (59 - 46)

Say 6.0

=

17.46 0 F

=

17.5 0 F

ESTIMATE NUMBER OF ROWS GTH x SHF Number of rows

= K x LMTD x Ws x FA 4,57,000 x 0.66 =

Say

=

173.20 x 17.5 x 0.891 x 29 3.85 rows

=

4 rows ISSUE R2

TCE CONSULTING ENGINEERS LIMITED GUIDE FOR SELECTION OF COOLING AND HEATING COILS

TCE.M6-ME-811-304

SECTION: APPENDIX SHEET 18 OF 49

APPENDIX-1 (CONTD.)

7.0

Water side pressure drop and air side pressure drop can be estimated from the water pressure drop and air pressure drop tables or charts.

8.0

SUMMARY OF PARAMETERS Coil face area, FA

= 2.7 M2 (29 FT2 )

Total cooling capacity

= 38 TR (4,57,000 BTU/Hr)

Chilled water flow rate, Qcw

= 20.5 M3 /Hr

Number of rows deep

= 4

Fin spacing

= 3.15 fins/cm (8 fins/inch)

EXAMPLE-3: DIRECT EXPANSION VOLATILE REFRIGERANT COILS 1.0

INPUT DATA Same as for example 1.

2.0

SELECT COIL FACE AREA From Table-2 select AH-9 with 22.6 FT2 (2.10 M2 ) coil face area . Air face velocity

= 14000 / 22.6 = 619.47 FPM

Say

= 625 FPM

This air face velocity is on a higher side of the recommended velocity of around 2.5 M/sec. However, the refrigerant evaporating temperature selection method is explained below. Grand Total Heat (GTH) 3.0

= 4,57,000 BTU/Hr

ESTIMATE REFRIGERANT EVAPORATING TEMPERATURE In Figure-9, enter at upper right hand side with GTH = 4,57,000 BTU/Hr and move upto intersection with 625 FPM curve. From this point project down to intersect with air entering wet bulb temperature curve of 68 o F, reading refrigerant evaporating temperature of approximately 50 o F. ISSUE R2

TCE CONSULTING ENGINEERS LIMITED TCE.M6-ME-811-304

SECTION: APPENDIX

GUIDE FOR SELECTION OF COOLING AND HEATING COILS

SHEET 19 OF 49

APPENDIX-1 (CONTD.)

4.0

SUMMARY OF PARAMETERS Coil face area, FA

= 2.10 M2 (22.6 FT2 )

Total cooling load, GTH

= 38 TR (4,57,000 BTU/Hr)

Refrigerant (R - 22) Refrigerant evaporating temperature = 10 o C (50 o F) Number of rows deep

=5

(from Figure-9)

Fin spacing

= 5 fins/cm (13 fins/inch)(from Figure-9)

Air side pressure drop

= 0.89 INCHES WC ( 23mmWC) (from Table-3)

ISSUE R2

TCE CONSULTING ENGINEERS LIMITED GUIDE FOR SELECTION OF COOLING AND HEATING COILS

TCE.M6-ME-811-304

SECTION: APPENDIX SHEET 20 OF 49

APPENDIX-2 EXAMPLES OF HEATING COIL SELECTION PROCEDURE For abbreviations and suffixes, refer Appendix-1. EXAMPLE-1: HOT WATER COIL - BASED ON RATING CHARTS AND TABLES FURNISHED BY BLUE STAR LIMITED

1.0

2.0

INPUT DATA (a)

Heating load

= 1,02,000 BTU/Hr

(b)

Air flow rate, Qa

= 2,500 CFM

(c)

Entering air DBT , tadb1

= 60 0 F

(d)

Air face velocity, Vaf

= 600 FPM

(e)

Water pressure drop, ∆phw

= 10 feet of water

(f)

Entering water temperature , thw1

= 150 0 F

(g)

Leaving water temperature, thw2

= 138 0 F

DETERMINE COIL FACE AREA Qa Coil face area =

= 4.17 FT2

= Vaf

3.0

2,500 600

TUBES HIGH (TH) AND FINNED LENGTH (FL) From Table-7 select TH =10 and FL =48 INCHES Coil face area ,FA = 4.2 FT2 Several combinations of TH and FL are possible to obtain same area. Select TH and FL based on layout consideration.

4.0

KBH / FT2 , where KBH = 1,000 BTU/Hr 102 = 24.5 KBH/ FT2 4.17 ISSUE R2

TCE CONSULTING ENGINEERS LIMITED GUIDE FOR SELECTION OF COOLING AND HEATING COILS

TCE.M6-ME-811-304

SECTION: APPENDIX SHEET 21 OF 49

APPENDIX-2 (CONTD.) 5.0

CALCULATE QB QB

= (Average water temp – tadb1 ) x 0.308 (constant) = (144-60) x 0.308 = 25.9 KBH/ FT2

6.0

For heating, F1 (water temperature rise factor) = 1.0

7.0

ESTIMATE HOT WATER FLOW RATE KBH x 2 Hot water flow rate, Qhw =

=

∆thw

= 17 USGPM 12

= 3.9 M3 /Hr

= 17/4.4 8.0

102 x 2

ESTIMATE HOT WATER PRESSURE DROP In Figure-10 choose a water circuit, taking into consideration any limitation on Water Pressure Drop (WPD) Accordingly choose circuit B, WPD = 6 FT Qhw (USGPM)x R Calculate water velocity, Vhw

= TH x RD

Where R= water velocity constant to be obtained from Table-8 = 6.84 TH= Tubes High

=10 (from para 3.0)

RD= Rows Deep

= 2 (assumed)

17 x 6.84 Vhw =

= 5.81 FPS 10 x 2

Vhw of 5.81 FPS is within acceptable limit. Now correct the 6 FT WPD for Vhw of 5.81 From Figure-10 multiplication factor = 1.95 6 FT x 1.95 = 11.7 FT Now correct for thw1 of 150 0 F

ISSUE R2

TCE CONSULTING ENGINEERS LIMITED GUIDE FOR SELECTION OF COOLING AND HEATING COILS

TCE.M6-ME-811-304

SECTION: APPENDIX SHEET 22 OF 49

APPENDIX-2 (CONTD.) From formula given in Figure-10 Heating WPD = WPD[0.75 +0.0012 (180-thw1 )] = 11.7 x [ 0.75 + 0.0012 (180-150)] = 9.2 FT WPD of 9.2 FT is within specified limit. 9.0

OBTAIN WATER VELOCITY FACTOR , F2 From Figure-11 read F2 = 1.0

10.0

CALCULATE F3 KBH / FT2 required = QB x F1 x F2 x F3 KBH / FT2 = 24.5 (from para 4.0) ; F1 = 1.0 (from para 6.0)

Where,

F2 = 1.0 (from para 9.0) 24.5 = 25.9 x 1 x 1 x F3 which gives 11.0

F3 = 0.95

SELECT FINS PER INCH (FPI) In Figure-12, for 2 rows deep, F3 = 0.95 and air face velocity = 600 FPM, select a FPI that provides a value of F3 greater than required. Choose 10 FPI. Then F3 = 1.0

12.0

ESTIMATE ACTUAL KBH Actual KBH / FT2

= QB x F1 x F2 xF3 = 25.9 x 1 x 1 x 1

Actual KBH

= 25.9

= 25.9 x 4.17 = 108 KBH or 1,08,000 BTU/Hr

13.0

CALCULATE LEAVING AIR DRY BULB TEMPERATURE (tadb2 ) KBH x 1000 tadb2

=

tadb1 +

108.0 x 1,000 = 100 0 F

= 60 + 1.08 x Qa

1.08 x 2,500 ISSUE R2

TCE CONSULTING ENGINEERS LIMITED GUIDE FOR SELECTION OF COOLING AND HEATING COILS

TCE.M6-ME-811-304

SECTION: APPENDIX SHEET 23 OF 49

APPENDIX-2 (CONTD.) 14.0

CALCULATE AIR PRESSURE DROP (∆pa) From Figure-13 , Y = 0.082 Air pressure drop, ∆pa = Y x Rows deep = 0.082 x 2 = 0.164 INCHES WC

15.0

= 4.17 mmWC

SUMMARY OF PARAMETERS Coil face area, FA

= 0.39M2 ( 4.17 FT2 )

Total heating capacity

= 25,760 Kcals/Hr (1,02,000 BTU/Hr)

Hot water flow rate, Qhw

= 3.9 M3 /Hr

Number of rows deep

= 2

Fins spacing

= 4 fins /cm (10 fin/inch)

Water side pressure drop, ∆pw= 2.8MWC Air side pressure drop, ∆pa

= 4.17 mmWC

EXAMPLE-2: HOT WATER COIL - BASED ON RATING CHARTS AND TABLES FURNISHED BY VOLTAS LIMITED 1.0

INPUT DATA (a)

Outside design temperature, to

= 0 0F

(b)

Inside design temperature, ti

= 70 0 F

(c)

Heating load

= 2,00,000 BTU/Hr

(d)

Total air flow rate, Qa

= 4,000 CFM

(e)

Outside air flow rate, Qo

= 800 CFM

(f)

Entering water temperature, thw1

= 180 0 F

(g)

Unit size(Selected based on total cooling load, dehumidified CFM and air face velocity)

= AH-5

Air face velocity, Va

= 500 FPM

(h)

ISSUE R2

TCE CONSULTING ENGINEERS LIMITED GUIDE FOR SELECTION OF COOLING AND HEATING COILS

TCE.M6-ME-811-304

SECTION: APPENDIX SHEET 24 OF 49

APPENDIX-2 (CONTD.) 2.0

Calculate the temperature of the air mixture entering the coil Qo x to + Qi x ti tmixture = Qa (800 x 0 ) + (3,200 x 70) = 56 0 F

= 4,000 3.0

DETERMINE BTU CONSTANT From Table-9 for water temperature drop

= 20 0 F,

entering water temperature

= 180 0 F and

entering air temperature

= 56 0 F,

interpolating between 50 0 F and 60 0 F , the hot water BTU constant = 0.87 4.0

CALCULATE TABLE BTU/HR Dividing the given heating load = 2,00,000 BTU/Hr by the BTU constant = 0.87, Table BTU/Hr = 2,30,000 BTU/Hr This value is to be used in Table-10 which is based on entering water temperature = 200 0 F, entering air temperature =60 0 F and hot water temperature drop =20 0 F.

5.0

DETERMINE HOT WATER FLOW RATE AND PRESSURE DROP From Table-10 for 2 row coil, unit size = AH-5 and face velocity = 500 FPM the BTU/Hr is 2,48,000, Qhw = 25 USGPM = 25 x 0.227 M3 /Hr

= 5.7 M3 /Hr

∆phw = 0.98 psi

= 0.68 MWC

=0.98 x 0.69 MWC

ISSUE R2

TCE CONSULTING ENGINEERS LIMITED GUIDE FOR SELECTION OF COOLING AND HEATING COILS

TCE.M6-ME-811-304

SECTION: APPENDIX SHEET 25 OF 49

APPENDIX-2 (CONTD.) 6.0

FIND THE LEAVING AIR TEMPERATURE BTU / Hr required Tadb2

= tadb1 + 1.08 x CFM 2,00,000 = 102.3 0 F

= 56 + 1.08 x 4,000 7.0

DETERMINE ACTUAL TEMPERATURE DROP BTU/Hr required

Water temperature drop, ∆thw =

500 x Water Flow Rate (USGPM) 2,00,000 = 16 0 F

= 500 x 25 8.0

SUMMARY OF PARAMETERS Coil face area, FA

= 0.767 M2 (from Table-2)

Total heating capacity

= 50,500 Kcal/Hr

Hot water flow rate, Qhw

= 5.7 M3 /Hr.

Number of rows deep

= 2

(2,00,000 BTU/Hr)

Water side pressure drop,∆phw= 0.68 MWC Air side pressure drop , ∆pa

= 3 mmWC (0.12 INCHES WC) (from Table-3)

EXAMPLE-3: STEAM HEATING COIL : BASED ON RATING CHARTS AND TABLES FURNISHED BY VOLTAS LIMITED

1.0

INPUT DATA (a)

Outside design condition, to = 0 0 F

(b)

Inside design condition, ti

= 70 0 F

(c)

Heating Load

= 2,80,000 BTU/Hr ISSUE R2

TCE CONSULTING ENGINEERS LIMITED GUIDE FOR SELECTION OF COOLING AND HEATING COILS

TCE.M6-ME-811-304

SECTION: APPENDIX SHEET 26 OF 49

APPENDIX-2 (CONTD.)

(d)

Total air flow rate, Qa

= 4,000 CFM

(e)

Outside air flow rate,Qo

= 800 CFM

(f)

Unit size ( selected based on total cooling load, dehumidified CFM and air face velocity)

2.0

(g)

Air face velocity ,Va

= 500 FPM

(h)

Steam pressure available

= 10 psig

= AH-5

Calculate the temperature of the mixed air entering the coil using the formula Qo x to + Qi x ti tmixture = Qa (800 x 0) + (3,200 x70) = 56 0 F

= 4,000 3.0

ESTIMATE EQUIVALENT HEATING COIL CAPACITY As 10 psig steam is available, calculate the equivalent capacity of a heating coil when operating at 2 psig steam and 70 0 F entering air 10 psig saturated steam temperature = 240 0 F BTU/Hr required x 149 Equivalent capacity =

(Voltas Handbook) Steam temp. - entering air temp. 2,80,000x 149

=

= 2,26,740 BTU/Hr 240 - 56

Say 4.0

= 2,27,000 BTU/Hr

SELECT THE COIL From Table-11 for AH-5 heating coil, for 2 psig steam and entering air temperature = 70 0 F and coil air face velocity = 500 FPM, select a 1 row high rise coil. Capacity = 2,33,000 BTU/Hr. This capacity will work satisfactorily. ISSUE R2

TCE CONSULTING ENGINEERS LIMITED GUIDE FOR SELECTION OF COOLING AND HEATING COILS

TCE.M6-ME-811-304

SECTION: APPENDIX SHEET 27 OF 49

APPENDIX-2 (CONTD.)

The actual capacity of the coil = 2,33,000 x (240 - 56) / 149 = 2,88,000 BTU/Hr 5.0

DETERMINE LEAVING AIR TEMPERATURE The leaving air temperature is found from the formula : BTU/Hr required tadb2

=

tadb1

+ 1.08 x Qa

2,80,000 = 120.8 0 F

= 56 + 1.08 x 4,000 6.0

SUMMARY OF PARAMETERS Coil face area

= 0.767 M2 (from Table-2)

Total heating capacity

= 70,700 Kcal/Hr (2,80,000 BTU/Hr)

Number of rows deep

= 1

Air side pressure drop , ∆pa

= 2 mmWC (0.08 INCHES WC) (from Table-3)

ISSUE R2

TCE CONSULTING ENGINEERS LIMITED

SECTION: TABLE

GUIDE FOR SELECTION OF COOLING AND HEATING COILS

SHEET 39 OF 49

TCE.M6-ME-811-304

TABLE - 1 RANGES OF STANDARD RATING CONDITIONS SL. NO.

1. 2. 3. 4.

5. 6.

7.

8. 9. 10. 11. 12.

ITEM

UNIT VOLATILE REFRIGERANT

Standard air face velocity Entering air drybulb temperature Entering air wetbulb temperature Tube-side fluid velocity Entering fluid temperature Saturated suction refrigerant temperature at coil outlet Minimum suction vapour superheat at coil outlet Steam pressure at coil inlet Maximum superheat in steam at coil inlet Concentration by mass Minimum fin surface temperature Minimum tube wall surface temperature

M/ sec 0

COOLING COILS COLD COLD WATER ETHYLENE GLYCOL SOLUTION

STEAM

HEATING COILS HOT HOT WATER ETHYLENE GLYCOL SOLUTION

1 to 4

1 to 4

1 to 4

1 to 8

1 to 8

1 to 8

18 to 38

18 to 38

18 to 38

16 to 29

16 to 29

16 to 29

(-) 29 to 38 --

(-) 18 to 38 --

(-) 29 to 38 --

--

0.3 to 2.4 (N1)

0.3 to 1.8 (N2)

--

0.1 to 1.8 (N2)

--

1.7 to 18

(-) 18 to 32

--

0.1 to 2.4 (N1) 49 to 121

C

0

C M/ Sec

(-) 18 to 93

0

C

0

C

(-) 1.1 to 13

--

--

--

--

--

0

C

3.3

--

--

--

--

--

KPa (g)

--

--

--

--

--

--

--

--

14 to 1723 28

--

--

--

--

10 to 60

--

--

10 to 60

> 0.0

> 0.0

> 0.0

> 0.0

> 0.0

> 0.0

0

C

% 0

C

0

C

> 0.0

> 0.0

> ethylene glycol solution freeze point

> 0.0

> 0.0

> ethylene glycol solution freeze point

NOTES 1.

On lower limit, Reynolds Number shall exceed 3,100 at mean water temperature.

2.

On lower limit, Reynolds Number shall exceed 700 at mean glycol temperature.

SOURCE: ARI 410-2001

ISSUE R2

DX

500 x 500 x 50

625 x 400 x 50

625 x 500 x 50

625 x 375 x 25

625 x 400 x 25

2 ROWS MINIMUM LOAD 3,4 5 ROWS FOR&DX COILS SOURCE : VOLTAS HANDBOOK-1977

MINIMUM LOAD FOR COILS (1000 Kcals/Hr)

FILTERS

MAXIMUM OPERATING SPEED RPM

1

1800

0.048

190

22.119

6 ROW

WHEEL SIZE - DIAMETER , mm

18.401

5 ROW

1

14.684

4 ROW

NUMBER OF FANS AND OUTLETS

11.060

3 ROW

570

1

7.342

850

1.5

2 ROW

2515

1140

2

AH-3

3110

4150

5185

6220

7260

1

1500

0.079

240

1

34.108

28.439

22.700

17.100

10.409

0.236 10

4.55 9.10

6.06

2

1100

0.165

325

1

84.572

70.167

56.227

42.286

27.881

0.567 16

3.03

2

1100

0.117

325

1

56.227

46.933

37.639

28.067

18.587

0.344 10

COOLING AND HEATING COILS

865

1890

3160

2.5

1290

3790

1715

4420

3

OUTSIDE SURFACE AREA, M2

OUTLET AREA, M

AH-2

3.5

0.155 8

2

3

AH-1

AH-5

AH-6

18.20

9.10

6

1100

0.242

380

1

123.606

102.695

82.249

61.803

40.892

0.767 20

4210

5610

7015

8415

9820

18.20

9.10

6

900

0.325

457

1

138.011 1 166.357

110.595

83.178

55.298

1.041 20

5715

7615

9520

11425

13330

CAPACITY IN M /HR BASED ON COIL FACE VELOCITIES

AH-0

COIL FACE AREA , M NUMBER OF TUBES IN FACE

2

AIR FACE VELOCITY, M/sec

UNIT SIZE

36.40

18.20

6

4

1100

0.483

380

2

333.64

206.32

165.43

124.07

82.714

1.626 20

8925

11900

14875

17850

20825

AH-7

TABLE-2 PHYSICAL DATA - AIR HANDLING UNITS

36.40

18.20

12

900

0.651

457

2

248.14

277.89

222.12

166.82

111.06

2.100 20

11525

15370

19200

23050

26900

AH-9

54.6

27.3

18

900

0.976

457

3

500.00

416.36

333.64

250.00

166.82

3.160 30

17350

23120

28900

34680

40460

AH-11

72.8

36.4

24

900

0.976

457

3

667.29

555.76

444.24

333.64

222.12

4.201 40

23050

30735

38420

46100

53790

AH-12

TCE.M6-ME-811-304

TCE CONSULTING ENGINEERS LIMITED SECTION: TABLE

GUIDE FOR SELECTION OF COOLING AND HEATING COILS SHEET 40 OF 49

ISSUE R2

TCE CONSULTING ENGINEERS LIMITED

SECTION: TABLE

GUIDE FOR SELECTION OF COOLING AND HEATING COILS

SHEET 41 OF 49

TCE.M6-ME-811-304

TABLE-3 AIR SIDE RESISTANCE OF COOLING AND HEATING COILS (INCHES OF WATER COLUMN) SL. NO.

AIR FACE VELOCITY FPM

2 ROW

COOLING COIL (NOTE 2) 3 4 5 ROW ROW ROW

HEATING COIL 1 ROW 2 LOW HIGH ROW RISE RISE

6 ROW

(NOTE 1)

1. 2. 3. 4. 5. 6. 7. 8. 9.

300 350 400 450 500 550 600 650 700

0.12 0.15 0.18 0.22 0.26 0.30 0.34 0.37 0.40

0.18 0.23 0.28 0.33 0.39 0.45 0.51 0.55 0.66

0.24 0.30 0.37 0.44 0.52 0.60 0.68 0.74 0.81

0.30 0.38 0.47 0.55 0.65 0.76 0.85 0.93 1.01

0.36 0.45 0.55 0.66 0.78 0.90 1.00 1.10 1.21

0.01 0.01 0.02 0.02 0.03 0.03 0.04 0.04 0.05

0.03 0.04 0.06 0.07 0.08 0.10 0.11 0.13 0.15

0.05 0.07 0.08 0.10 0.12 0.14 0.17 0.19 0.22

NOTES 1.

Air quantity for each unit size corresponding to the air face velocity is shown in the physical data in Table-2.

2.

Cooling coil resistance is based on wet surface.

SOURCE : VOLTAS HANDBOOK-1977

ISSUE R2

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SECTION: TABLE

GUIDE FOR SELECTION OF COOLING AND HEATING COILS

SHEET 42 OF 49

TABLE 4 PHYSICAL PARAMETERS OF COILS ( 8 FINS/INCH) MODEL NO.

COILS

PIPES HIGH

CONNECTION SIZES DX COILS WATER COILS

FAH-12

10

NOMINAL FACE LENGTH INCHES 22

FACE AREA

INLET OD

SUCTION OD

INLET OD

OUTLET OD

FT2 2.5

INCHES 5/8

INCHES 7/8

INCHES 7/8

INCHES 7/8

FAH-21

12

34

4.2

3/4

1-1/8

1-1/8

1-1/8

FAH-32

14

46

6.4

3/4

1-3/8

1-3/8

1-3/8

FAH-38

16

46

7.7

7/8

1-5/8

1-5/8

1-5/8

FAH-48

16

56

9.7

7/8

1-5/8

1-5/8

1-5/8

FAH-65

16

78

13.0

7/8

1-5/8

1-5/8

1-5/8

FAH-86

18

94

17.2

1-1/8

2-1/8

2-1/8

2-1/8

FAH-98

24

78

19.6

1-1/8

2-1/8

2-1/8

2-1/8

FAH-115

24

94

23.6

1-1/8

2-1/8

2-1/8

2-1/8

FAH-145

28

102

29.0

2 x 7/8

FAH-170

28

118

34.4

FAH-230

38

118

46.7

2 x 11/8 2 x 11/8

2x 1-5/8 2x 2-1/8 2x 2-1/8

2 x 15/8 2 x 21/8 2x 21/8

2x 1-5/8 2x 2-1/8 2x 2-1/8

SOURCE : FRICK INDIA LIMITED – ENGINEERING BULLETIN

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TCE.M6-ME-811-304

TABLE-5 K FACTORS FOR CHILLED WATER COILS (8 FINS/INCH) BTU PER HOUR PER ROW (SENSIBLE HEAT ) PER SQ.FT FACE AREA PER DEG. F LMTD

SL. NO.

AIR FACE VELOCITY FPM

1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14. 15.

350 375 400 425 450 475 500 525 550 575 600 625 650 675 700

WATER VELOCITY IN COIL – FEET PER SECOND 1 110 115 118 122 125 128 132 135 138 140 143 146 148 150 152

1.5 119 124 129 144 148 152 156 160 164 168 172 176 179 182 186

2 129 134 139 144 148 152 156 160 164 168 172 176 179 182 186

2.5 132 138 143 148 153 157 161 166 170 174 179 184 188 191 195

3 135 141 147 153 154 164 169 173 178 183 187 192 190 200 204

3.5 138 144 150 156 162 168 173 178 183 187 192 196 201 205 210

4 141 147 154 160 166 172 178 183 188 194 198 203 208 212 217

5 143 150 156 163 169 175 182 187 192 197 203 208 212 217 222

6 145 153 159 166 172 179 185 190 196 204 207 212 217 222 227

SOURCE : FRICK INDIA LIMITED – ENGINEERING BULLETIN

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TCE.M6-ME-811-304

TABLE-6 MULTIPLIERS WS FOR WETTED CHILLED WATER COOLING COIL SURFACES (8 FINS PER INCH)

SL. NO. 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14. 15. 16. 17. 18. 19. 20. 21. 22. 23. 24. 25. 26.

SENSIBLE HEAT FACTOR 0.40 0.42 0.44 0.46 0.48 0.50 0.52 0.54 0.56 0.58 0.60 0.62 0.64 0.66 0.68 0.70 0.72 0.74 0.76 0.78 0.80 0.82 0.84 0.86 0.88 0.90

1 0.605 0.624 0.643 0.662 0.680 0.698 0.713 0.728 0.743 0.759 0.772 0.785 0.800 0.813 0.826 0.839 0.850 0.864 0.875 0.888 0.899 0.910 0.921 0.933 0.944 0.955

WATER VELOCITY (Vcw) IN COIL ( FEET PER SECOND) 1.5 2 2.5 3 3.5 4 5 0.638 0.657 0.676 0.695 0.714 0.728 0.740 0.757 0.770 0.782 0.796 0.808 0.822 0.835 0.848 0.860 0.871 0.880 0.890 0.901 0.911 0.921 0.931 0.941 0.951 0.960

0.688 0.687 0.706 0.724 0.739 0.754 0.769 0.783 0.796 0.810 0.823 0.832 0.846 0.855 0.867 0.878 0.888 0.899 0.907 0.916 0.925 0.933 0.942 0.950 0.956 0.966

0.690 0.705 0.720 0.735 0.752 0.766 0.782 0.796 0.810 0.822 0.835 0.847 0.857 0.867 0.876 0.886 0.896 0.905 0.914 0.923 0.932 0.939 0.945 0.952 0.959 0.967

0.707 0.723 0.739 0.755 0.772 0.785 0.800 0.813 0.824 0.837 0.849 0.859 0.868 0.877 0.887 0.897 0.906 0.913 0.922 0.930 0.938 0.942 0.950 0.957 0.963 0.970

0.722 0.737 0.750 0.767 0.784 0.798 0.811 0.822 0.833 0.844 0.855 0.866 0.876 0.885 0.893 0.902 0.911 0.919 0.926 0.933 0.940 0.946 0.952 0.960 0.964 0.968

0.736 0.750 0.764 0.780 0.795 0.807 0.820 0.831 0.841 0.851 0.862 0.871 0.881 0.891 0.900 0.908 0.916 0.924 0.931 0.938 0.944 0.951 0.957 0.963 0.970 0.975

0.750 0.763 0.777 0.790 0.803 0.816 0.829 0.842 0.853 0.861 0.872 0.880 0.889 0.898 0.907 0.915 0.923 0.929 0.936 0.942 0.948 0.954 0.960 0.965 0.971 0.976

6 0.765 0.770 0.787 0.803 0.818 0.829 0.841 0.852 0.860 0.868 0.880 0.888 0.896 0.905 0.913 0.921 0.928 0.935 0.941 0.947 0.952 0.957 0.962 0.967 0.972 0.977

SOURCE : FRICK INDIA LIMITED – ENGINEERING BULLETIN

ISSUE R2

10.0

0.8 1.3 1.7 2.1 2.5 2.9 3.3

7.5

0.6 0.9 1.3 1.6 1.9

FINNED HEIGHT (FH) - INCHES 12 18 24 30 36 42 48 54 60 66 72 78 84 90 96 102 108 114 120 126 132 140

FINNED LENGTH (FL) - INCHES

1.9 2.5 3.1 3.8 4.4 5.0 5.6 6.3 6.9 7.5

15

12

2.2 2.9 3.7 4.4 5.1 5.8 6.6 7.3 8.0 8.8 9.5 10.2

17.5

14

3.3 4.2 5.0 5.8 6.7 7.5 8.3 9.2 10.0 10.8 11.7 12.5 13.3

20.0

3.8 4.7 5.6 6.6 7.5 8.4 9.4 10.3 11.3 12.2 13.1 14.1 15.0 15.9 16.9

22.5

5.2 6.3 7.3 8.3 9.4 10.4 11.5 12.5 13.5 14.6 15.6 16.7 17.7 18.8 19.8 20.8 21.9 22.9 24.0

25.0

5.7 5.9 8.0 9.2 10.3 11.5 12.6 13.8 14.9 16.0 17.2 18.3 19.5 20.6 21.8 22.9 24.1 25.2 26.3

27.5

22

7.5 8.8 10.0 11.3 12.5 13.8 15.0 16.3 17.5 18.8 20.0 21.3 22.5 23.8 25.0 26.3 27.5 28.8

30.0

24

8.1 9.5 10.8 12.2 13.5 14.9 16.9 17.6 19.0 20.3 21.7 23.0 24.4 25.7 27.1 28.4 29.7 31.0

32.5

26

10.2 11.7 13.1 14.6 16.0 17.5 19.0 20.4 21.9 23.3 24.8 26.3 27.7 29.2 30.6 32.0 33.5

35.0

28

10.9 12.5 14.1 15.6 17.2 18.8 20.3 21.9 23.4 25.0 26.6 28.1 29.7 31.3 32.8 34.4 35.9

37.5

30

13.3 15.5 16.7 18.3 20.0 21.7 23.3 25.0 26.7 28.3 30.0 31.7 33.3 35.0 36.7 38.4

40.0

32

FINNED LENGTH AVAILABLE IN 2 INCH INCREMENTS FROM 12 TO 140 INCHES

1.0 1.6 2.1 2.6 3.1 3.7 4.2 4.7 5.2

12.5

10

SOURCE: BLUE STAR LIMITED – ENGINEERING BULLETIN

8

6

TUBES HIGH (TH)

(SQUARE FEET) 16 18 20

TABLE-7 COIL FACE AREAS (FA)

14.2 15.9 17.7 19.5 21.3 23.0 24.8 26.6 28.3 30.1 31.9 33.7 35.4 37.2 38.9 40.7

42.5

34

16.9 18.8 20.6 22.5 24.4 26.3 28.1 30.0 31.9 33.8 35.6 37.5 39.4 41.3 43.2

45.0

36

TCE.M6-ME-811-304

TCE CONSULTING ENGINEERS LIMITED SECTION: TABLE

GUIDE FOR SELECTION OF COOLING AND HEATING COILS SHEET 45 OF 49

ISSUE R2

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GUIDE FOR SELECTION OF COOLING AND HEATING COILS

SHEET 46 OF 49

TABLE-8 WATER VELOCITY CONSTANT ‘R’ COIL CIRCUIT

COOLING ROWS DEEP AVAILABLE

HEATING ROWS DEEP SAVAILABLE

A

3,4

1,2,3,4

B

3,4,5,6,8

1,2,3,4

C

3,4,5,6,8

1,2,3,4

D

4,5,6,8

4

E

4,5,6,8

4

F

4,5,6,8

4

TUBE WALL THICKNESS INCHES 0.165(STD) 0.020 0.032 0.165(STD) 0.020 0.032 0.165(STD) 0.020 0.032 0.165(STD) 0.020 0.032 0.165(STD) 0.020 0.032 0.165(STD) 0.020 0.032

CONSTANT R 3.42 3.52 3.89 6.84 7.04 7.78 10.26 10.56 11.67 13.68 14.08 15.56 20.52 21.12 23.34 27.36 28.16 31.12

ALL COILS AVAILABLE WITH FIN SPACINGS OF 8,10 OR 12 FINS PER INCH

SOURCE: BLUE STAR LIMITED - ENGINEERING BULLETIN

ISSUE R2

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GUIDE FOR SELECTION OF COOLING AND HEATING COILS

SHEET 47 OF 49

TABLE-9 HOT WATER BTU CONSTANTS (BASED ON 200 0 F ENTERING WATER, 20 0 F WATER TEMPERATURE DROP THROUGH UNIT AND 60 0 F ENTERING AIR TEMPERATURE) WATER TEMPERATURE DROP THROUGH UNIT 0 F

20

30

40

50

TEMPERATURE OF ENTERING AIR

ENTERING WATER TEMPERATURE , 0 F 140

160

180

200

220

240

260

280

0.69 0.59 0.48 0.40 0.33 0.63 0.56 0.47 0.38 0.30 0.57 0.49 0.40 0.32 0.22 0.48 0.41 0.32 0.23 0.15

0.86 0.77 0.66 0.58 0.50 0.78 0.70 0.62 0.53 0.45 0.71 0.63 0.55 0.47 0.37 0.63 0.56 0.47 0.38 0.30

1.02 0.93 0.83 0.75 0.67 0.92 0.85 0.77 0.68 0.59 0.86 0.77 0.69 0.61 0.52 0.78 0.70 0.62 0.53 0.45

1.19 1.10 1.00 0.91 0.83 1.07 0.99 0.91 0.82 0.72 1.00 0.92 0.83 0.75 0.67 0.92 0.85 0.77 0.68 0.59

1.35 1.26 1.16 1.07 0.99 1.22 1.13 1.06 0.97 0.88 1.15 1.07 0.99 0.90 0.82 1.07 0.99 0.91 0.82 0.72

1.50 1.41 1.32 1.22 1.13 1.36 1.28 1.21 1.12 1.03 1.30 1.22 1.14 1.04 0.96 1.22 1.13 1.06 0.97 0.88

1.65 1.56 1.47 1.37 1.29 1.53 1.44 1.35 1.27 1.17 1.48 1.38 1.29 1.19 1.10 1.36 1.28 1.21 1.12 1.03

1.78 1.70 1.62 1.52 1.42 1.70 1.62 1.52 1.43 1.34 1.63 1.54 1.47 1.33 1.25 1.53 1.44 1.35 1.27 1.17

0

F 40 50 60 70 80 40 50 60 70 80 40 50 60 70 80 40 50 60 70 80

SOURCE: VOLTAS HANDBOOK-1977

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TCE.M6-ME-811-304

TABLE-10 2 ROW HOT WATER COIL CAPACITIES (BASED ON 200 0 F ENTERING WATER, 20 0 F WATER TEMPERATURE DROP THROUGH UNIT AND 60 0 F ENTERING AIR ) UNIT SIZE

AH-2

AH-3

AH-5

AH-6

COIL FACE VELOCITY FPM 300 350 400 450 500 550 600 650 700 300 350 400 450 500 550 600 650 700 300 350 400 450 500 550 600 650 700 300 350 400 450 500 550 600 650 700

1000 BTU/HR

81 90 98 106 115 120 127 133 139 122 134 147 159 172 181 191 199 208 176 196 216 232 248 263 279 294 308 248 277 305 326 350 370 389 410 430

GPM

PRESSURE DROP PSI

8.7 9.6 10.6 11.5 12.3 13.0 13.6 14.2 14.9 13.0 14.4 15.9 17.2 18.4 19.5 20.4 21.3 22.4 10.0 20.0 21.5 23.5 25.0 27.0 28.5 30.0 31.5 25.0 28.0 31.0 33.0 35.0 37.0 39.0 41.0 43.0

0.52 0.60 0.72 0.85 0.97 1.06 1.16 1.25 1.38 0.51 0.62 0.75 0.87 1.00 1.11 1.21 1.33 1.45 0.52 0.63 0.72 0.87 0.98 1.13 1.24 1.36 1.49 0.98 1.21 1.47 1.65 1.82 2.04 2.30 2.52 2.73

UNIT SIZE

AH-7

AH-9

AH11

AH12

COIL FACE VELOCITY FPM 300 350 400 450 500 550 600 650 700 300 350 400 450 500 550 600 650 700 300 350 400 450 500 550 600 650 700 500 550 600 650 700

1000 BTU/HR

388 428 469 507 545 575 608 640 672 532 580 643 695 740 792 836 879 925 797 884 965 1042 1122 1188 1255 1318 1388 1495 1584 1672 1758 1850

GPM

39.5 43.0 47.0 51.0 54.5 58.0 60.5 64.0 67.0 53.0 59.0 64.0 69.0 75.0 79.0 83.0 88.0 92.0 79.5 88.5 96.5 103.5 112.5 118.5 124.5 132.0 138.0 114.0 158.0 165.0 175.0 184.0

PRESSURE DROP PSI 0.78 0.90 1.04 1.18 1.33 1.47 1.60 1.73 1.91 1.00 1.21 1.39 1.56 1.82 1.97 2.14 2.38 2.56 1.00 1.20 1.39 1.55 1.81 1.96 2.13 2.37 2.56 1.81 1.96 2.13 2.36 2.54

NOTE Capacity for conditions not above can be obtained by multiplying the BTU/Hr rating on the chart by the proper HOT WATER BTU CONSTANT (Table-9) BTU/Hr Temperature rise = 1.08 x CFM SOURCE :VOLTAS HANDBOOK-1977

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TCE.M6-ME-811-304

TABLE-11 STEAM HEATING COIL CAPACITIES (1 ROW LOW RISE AND HIGH RISE AND 2 ROW COILS) ENT. AIR TEMP. 0 F

STEAM PRESSURE PSI

70

2

35

2

70

2

35

2

70

2

35

2

70

2

35

2

ROWS OF HEAT -ING COIL 1 LR 1 HR 2 1 LR 1 HR 2 1 LR 1 HR 2 1 LR 1 HR 2 1 LR 1 HR 2 1 LR 1 HR 2 1 LR 1 HR 2 1 LR 1 HR 2

UNIT SIZE

AH2

AH3

AH5

AH6

COIL AIR FACE VELOCITY FPM 400 500 600 700 1000 BTU/HR 44 97 122 54.4 119 151 66 145 181 81.5 178 224 96 211 262 117 259 325 130 285 354 160 354 430

50.6 106 143 62.5 131 175 75 160 215 93.2 196 259 109.2 233 307 135.5 185 376 147 314 413 182.5 384 507

56.5 105 157 69.8 142 195 84.8 171 232 105.2 214 287 123.5 249 337 153 311 420 166.5 336 455 206.5 420 567

62.3 121 171 77 149 214 92.3 183 254 114 207 316 136 266 369 168 332 460 183.5 358 498 226.5 447 620

UNIT SIZE

AH7

AH9

AH11

AH12

COIL AIR FACE VELOCITY FPM 400 500 600 700 1000 BTU/HR 190 413 532 234.5 512 859 252.5 547 716 312 678 886 385 843 1068 476 1040 1321 505 1094 1432 624 1356 1772

217.5 456 622 275 568 762 292 602 837 370 743 1025 440 932 1250 548 1144 1532 584 1204 1674 740 1486 2050

244 494 683 300.5 609 852 333 652 922 406 807 1148 495.5 1013 1373 612 1245 1711 666 1304 1844 812 1614 2296

268 527 748 332 650 932 355 694 1008 440 858 1254 547 1059 1502 675 1305 1870 710 1388 2016 880 1716 2508

BTU/Hr Temperature Rise = 1.08 x CFM

SOURCE: VOLTAS HANDBOOK-1977

ISSUE R2

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SECTION: TITLE

TCE.M6-ME-811-307

DESIGN GUIDE FOR DUCT SIZING

SHEET i

OF iii

DESIGN GUIDE FOR DUCT SIZING

FILE NAME: M6ME307R2.DOC

REV. NO.

R0

R1

R2 ISSUE

INITIALS

SIGN.

INITIALS

SIGN.

INITIALS

PPD. BY

DSJ/BM/DJ

Sd/-/Sd/-/Sd/-

HRK

Sd/-

TPR

CHD. BY

SJB

Sd/-

PRJ

Sd/-

HRK

APD. BY

RRG

Sd/-

RL

Sd/-

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INITIALS

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R2

DATE

21.01.1986

15.02.2000

20.03.2003 TCE FORM NO. 020R2

TCE CONSULTING ENGINEERS LIMITED

SECTION: CONTENTS

DESIGN GUIDE FOR DUCT SIZING

SHEET ii OF iii

TCE.M6-ME-811-307

CONTENTS SL. NO.

TITLE

SH. NO.

1.0

SCOPE

1

2.0

INPUT DATA

1

3.0

METHODS OF DUCT SIZING

1

4.0

DUCT SIZING BY EQUAL FRICTION METHOD

2

5.0

REFERENCES

8

ISSUE R2 TCE FORM NO. 120 R1

TCE CONSULTING ENGINEERS LIMITED

SECTION: REV. STATUS

TCE.M6-ME-811-307

DESIGN GUIDE FOR DUCT SIZING

SHEET iii

OF iii

REVISION STATUS REV. NO.

DATE

DESCRIPTION

R0

21.01.1986

---

R1

15.02.2000

Completely revised.

R2

20.03.2003

Generally revised.

ISSUE R2 TCE FORM NO. 120 R1

TCE CONSULTING ENGINEERS LIMITED

SECTION: WRITE-UP

TCE.M6-ME-811-307

DESIGN GUIDE FOR DUCT SIZING

1.0

SHEET 1

OF

8

SCOPE This document gives guidelines for duct sizing for air-conditioning, ventilation and dust extraction systems.

2.0

INPUT DATA The following data shall be obtained from the preliminary single-line duct layout drawings:

3.0

(a)

Proposed duct routing with air flow rates in each individual duct branch

(b)

Type, sizes and locations of air outlets and inlets. Orientation of grilles to be indicated, i.e. location on side or bottom of duct.

(c)

Maximum allowable friction drop in the ducting system

(d)

Dust or powder loading on the duct. This item is required only for the dust extraction systems

METHODS OF DUCT SIZING The duct sizing is carried out by the following methods:

3.1

EQUAL FRICTION METHOD In this method the duct sizing is done for a constant pressure loss per unit length. This implies that the longest air flow path will dictate the maximum pressure drop in the system. This method is simple to use and made even easier by the use of ductulators.

3.2

STATIC REGAIN METHOD In this method the ducts for succeeding duct branches are sized so as to reduce the duct velocity and effect an increase in the air static pressure. This causes the duct sizes to be larger increasing the ducting system cost. This method is therefore, generally not used. For a more detailed description of this method of ducting system design, ASHRAE Handbook, Fundamentals, 2001 edition may be referred.

3.3

T-METHOD This is a computerised method, developed for optimisation of the ducting system design minimising the total costs considering the plant operation and installation costs. If required, ASHRAE Handbook, Fundamentals, 2001 edition may be referred.

ISSUE R2 TCE FORM NO. 120 R1

TCE CONSULTING ENGINEERS LIMITED

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DESIGN GUIDE FOR DUCT SIZING

SHEET 2

4.0

DUCT SIZING BY EQUAL FRICTION METHOD

4.1

The equal friction method is generally used in TCE for duct sizing. In this method the length of the longest branch governs the total static pressure drop in the system. In case the allowable pressure drop in the system is specified, the friction drop per unit length of the duct is given by the following expression: ∆p

=

OF

8

n

[∆P - Σ x ]] / L i i=1

Where, ∆p

=

Friction drop per unit length of duct (mmWG / M)

∆P

=

Allowable friction drop in the system (mmWG)

xi

=

Friction drop across elements in the system like cooling and heating coils, filters, dampers, supply and return grilles, etc. (mmWG)

n

=

Number of elements

L

=

Equivalent length of the duct (M).

The equivalent length of the duct shall be estimated for each of the fittings envisaged, from tables and chart listed below. These tables and charts are given in Part 2, Chapter 2, Handbook of Air-Conditioning System Design, Carrier Air-Conditioning Company, 1965 edition: (a)

Table 9

Friction of Round Duct System Elements

(b)

Table 10

Friction of Rectangular Duct System Elements

(c)

Table 11

Friction of Round Elbows

(d)

Table 12

Friction of Rectangular Elbows

(e)

Table 13

Percent Section Area in Branches for Maintaining Equal Friction

Losses for round fittings given in Chart 9 and the additional equivalent length of straight duct given in Table 11 and Table 12 listed above, are in FPS units and shall be converted to MKS units in the course of the calculations. The spacing and number of vanes for elbows shall be selected from Chart 6 Vane Location for Rectangular Elbows, Part 2, Chapter 2, Handbook of AirConditioning System Design, Carrier Air-Conditioning Company, 1965 edition.

ISSUE R2 TCE FORM NO. 120 R1

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DESIGN GUIDE FOR DUCT SIZING

SHEET 3

OF

4.2

DUCT SIZING CRITERIA

4.2.1

The equal friction method is normally used for duct sizing. As per Figure 9, Chapter 34, ASHRAE Handbook, Fundamentals, 2001 edition, for equal friction method, the friction drop is selected in the range of 0.06 to 0.5 mmWG per M of duct length or air velocity in duct between 9.0 to 20.0 M/sec, whichever is more stringent.

4.2.2

It is recommended that the ducts be sized for a pressure drop of 0.067 mmWG per M of duct length. The lower of the friction drop per unit length as calculated from equation given in para 4.1 or 0.067 mmWG per M of duct length as given above, shall be considered for the duct sizing.

4.2.3

The recommended air velocities for air-conditioning applications may be lower than the lower limit of 9.0 M/sec indicated in para 4.2.1 from criteria of noise reduction and application. Refer table 1 for recommendations on air velocity selection for various applications. In addition to the recommendations given in table 1, the recommended velocity for installations of indigenous packaged airconditioners considering the static pressures achievable is 7.5 M/sec. For ventilation systems the air velocity may be selected between 7.5 to 9.15 M/sec.

8

TABLE 1 RECOMMENDED MAXIMUM DUCT VELOCITIES FOR AIR-CONDITIONING AND VENTILATION SYSTEMS SL. NO.

APPLICATION

DUCT VELOCITY (M/sec) CONTROLLING FACTOR NOISE

MAIN DUCTS

BRANCH DUCTS

SUPPLY

RETURN

SUPPLY

RETURN

1.

RESIDENCES

3.0

5.0

4.0

3.0

3.0

2.

APARTMENTS, HOTEL BEDROOMS, HOSPITAL BEDROOMS

5.0

7.5

6.5

6.0

5.0

3.

PRIVATE OFFICES, DIRECTORS ROOMS, LIBRARIES

6.0

10.0

7.5

8.0

6.0

4.

THEATRES, AUDITORIUMS

4.0

6.5

5.5

5.0

4.0

5.

GENERAL OFFICES, HIGH CLASS RESTAURANTS, HIGH CLASS STORES, BANKS

7.5

10.0

7.5

8.0

6.0

6.

AVERAGE STORES, CAFETERIAS

9.0

10.0

7.5

8.0

6.0

7.

INDUSTRIAL

12.5

15.0

9.0

11.0

7.5

SOURCE: CARRIER, PART 2, CHAPTER 2., TABLE 7

ISSUE R2 TCE FORM NO. 120 R1

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DESIGN GUIDE FOR DUCT SIZING

SHEET 4

OF

8

TABLE 2 TRANSPORT VELOCITIES FOR DIFFERENT MATERIALS SL. NO.

MATERIAL

VELOCITY M/sec

FPM

SL. NO.

MATERIAL

VELOCITY M/sec

FPM

1.

Abrasive blasting

17.8 - 20.3

3500 - 4000

15.

Coal (powdered) dust

20.3

4000

2.

Aluminium coarse dust

20.3

4000

16.

Cocoa dust

15.2

3000

3.

Asbestos carding

15.2

3000

17.

Cork (ground) dust

12.7

2500

4.

Bakelite moulding powder dust

12.7

2500

18.

Cotton dust

15.2

3000

5.

Barrel filling or dumping

17.8-20.3

3500-4000

19.

Crushers

15.2 or higher

3000 or higher

6.

Belt conveyors

17.8

3500

20.

Flour dust

12.7

2500

7.

Bins and hoppers

17.8

3500

21.

Foundry, general

17.8

3500

8.

Brass turnings

20.3

4000

Sand mixer

17.8-20.3

3500-4000

9.

Bucket elevators

17.8

3500

Shakeout

17.8-20.3

3500-4000

10.

Buffing and polishing, Dry

15.2-17.8

3000-3500

Swing grinding booth exhaust

15.2

3000

11.

Buffing and polishing, Sticky

17.8-22.9

3500-4500

Tumbling mills

20.3-25.4

4000-5000

12.

Cast iron boring dust

22.9

4000

22.

Grain dust

12.7-15.2

2500-3000

13.

Ceramics, general

23.

Grinding general

17.8-22.9

3500-4500

14.

Glaze spraying

12.7

2500

24.

Portable hand grinding

17.8

3500

Brushing

17.8

3500

25.

Jute Dust

12.7-15.2

2500-3000

Fettling

17.8

3500

26.

Jute Lint

15.2

3000

Dry pan mixing

17.8

3500

27.

Jute dust shaker waste

16.3

3200

Dry press

17.8

3500

28.

Jute pickerstock

15.2

3000

Sagger filling

17.8

3500

29.

Lead dust

20.3

4000

Clay dust

17.8

3500

30.

Lead dust with small chips

25.4

5000

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8

TABLE 2 (CONTD.) 31.

Leather dust

17.8

3500

48

Soldering and tinning

12.7

2500

32.

Limestone dust

17.8

3500

49.

Spray painting

10.2

2000

33.

Lint

10.2

2000

50.

Starch dust

15.2

3000

34.

Magnesium dust, coarse

20.3

4000

51.

Stone cutting and finishing

17.8

3500

35.

Metal turnings

20.3-25.4

4000-5000

52.

Tobacco dust

17.8

3500

36.

Packaging, weighing, etc

15.2

3000

53.

37.

Downdraft grille

17.8

3500

Wood flour, light dry sawdust and shaving

12.7

2500

38.

Pharmaceutical coating pans

15.2

3000

Heavy shaving, damp sawdust

17.8

3500

39.

Plastic dust (buffing)

19.3

3800

Heavy wood chips waste, green shavings

20.3

4000

40.

Plating

10.2

2000

Hog waste

15.2

3000

41.

Rubber dust, Fine

12.7

2500

54.

Wool

15.2

3000

42.

Rubber dust, Coarse

22.9

4000

55.

Zinc oxide fume

10.2

2000

43.

Screens, Cylindrical

17.8

3500

56.

Very fine light dusts

10.2

2000

44.

Screens, Flat deck

17.8

3500

57.

Fine ,dry dusts and powders

15.2

3000

45.

Silica dust

17.8-22.9

3500-4500

58.

Average industrial dusts

17.8

3500

46.

Soap dust

15.2

3000

59.

Coarse dusts

20.3 - 22.9

4000 - 5000

47.

Soapstone dust

17.8

3500

60.

Heavy or moist dust loading

22.9 or higher

4500 or higher

Woodworking

SOURCE: ADVANCED DESIGN OF VENTILATION SYSTEMS FOR CONTAMINANT CONTROL BY HOWARD D. GOODFELLOW, ELSEVIER SCIENCE PUBLISHERS, B. V. 1985, TABLE 4.11, PAGE 320.

4.2.4

The duct velocities for dust extraction systems, however, may be much higher than the velocity range indicated in table 1, depending on the application. The recommended duct velocities for dust extraction systems are as given in table 2. However, for more details of dust or fume extraction system design, the designer may refer reference number 5.3 indicated at the end of this write-up. ISSUE R2 TCE FORM NO. 120 R1

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4.2.5

The duct sizing may be carried out using ductulators of Trane, Carrier or SMACNA make, or Chart 7 and Table 6, Part 2, Chapter 2, Friction Loss for Round Duct, Handbook of Air-Conditioning System Design, Carrier AirConditioning Company, 1965 edition. The chart may be used for systems handling air from (-)1 to 490 C (30 to 120 0 F) and for altitudes upto 600 M (2000 ft) without correcting for density. Air density correction factors for other conditions are given in Chart 15, Part 2, Chapter 2, of Handbook of AirConditioning System Design, Carrier Air-Conditioning Company, 1965 edition. However, correction factors for altitude and temperature for air density other than 0.075 lb/ft3 can be read directly using SMACNA make ductulators.

4.3

In case the duct sizing calculations are done manually, the friction drop in the ducting shall be calculated from following equation: ∆p

=

8

f x (100 x L x ρ x V2 ) 2 x Dh

Where, ∆p

=

Friction drop in unit length of duct (mmWG/ M)

f

=

Friction factor (Dimensionless)

L

=

Duct length (M)

ρ

=

Air density (Kg / M3 ) pa 2.871 x 105 x T

= pa

=

partial pressure of dry air (kPa)

T

=

Absolute temperature of air (0 K)

V

=

Air velocity (M/sec)

Dh

=

Hydraulic diameter of duct (mm)

=

[4 x Duct cross section area ( mm2 )] ÷ Duct perimeter (mm)

The values for air partial pressure shall be taken from Table 2, Chapter 6, ASHRAE Handbook, Fundamentals, 2001 edition. 4.4

The friction factor for the manual duct friction drop calculation shall be calculated from the following equation: 1.

[

∈ 3.7Dh

+

2.51 Re x f0.5

f0.5 Where,

=

(-)2 x log

f

=

Friction factor (Dimensionless)

] ISSUE R2 TCE FORM NO. 120 R1

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DESIGN GUIDE FOR DUCT SIZING

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∈

=

Material absolute roughness factor (mm). Refer Table 1, Chapter 34, ASHRAE Handbook, Fundamentals, 2001 edition, for duct roughness factors of various materials of construction.

Dh

=

Hydraulic diameter of duct (mm)

Re

=

Reynolds number

=

(Dh x V) ÷ (1000 x ν)

=

66.4 x Dh x V (For standard air)

V

=

Air velocity (M/sec)

ν

=

Kinematic viscosity of air (M2 /sec)

OF

8

The values for air viscosity shall be taken from Figure 12, Chapter 6, ASHRAE Handbook, Fundamentals, 2001 edition. 4.5

The equivalent rectangular duct diameter is given by the following equation: De

=

0.625 1.3 x (ab) 0.250 (a+b)

Where, De

=

Circular equivalent of rectangular duct for equal length, fluid resistance and air flow (mm)

a

=

Length of one side of duct (mm)

b

=

Length of adjacent side of duct (mm)

Normally, the value of the recommended maximum aspect ratio i.e. (a/b) or (b/a) is 4.0. However, this limitation may be ignored if the layout requires a flatter duct having an aspect ratio greater than 4.0. 4.6

For flat oval ducts the formula is modified as follows: De

A 0.625 P0.250

=

1.55 x

=

Cross sectional area of flat oval duct (mm2 )

=

(πb2 / 4) + b(a – b)

=

Perimeter (mm)

=

πb + 2(a - b)

=

Major dimension of oval duct (mm)

Where, A

P

a

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8

Minor dimension of oval duct (mm)

4.7

For the criteria given above the round duct diameter shall be obtained from the equation given in para 4.3. The sizes for the equivalent rectangular ducts shall be calculated from equation given in para 4.5. Equivalent duct sizes can also be directly read from Table 6, Part 2, Chapter 2, Handbook of Air-Conditioning System Design, Carrier Air-Conditioning Company, 1965 edition, or Table 2, Chapter 34, ASHRAE Handbook, Fundamentals, 2001 edition or from ductulators. These dimensions shall be decided considering the major criteria listed below. Refer TCE.M6-ME-811-318 “Design Guide for Duct Layout Drawings” for details.

4.7.1

Clear height requirements in the room

4.7.2

Clearances from structures, piping, cable trays or ducts

4.7.3

Access to other facilities

4.7.4

Sizes of equipment to which ducts are connected

4.7.5

Requirements of duct internal lining. The duct size to be enhanced to the extent of the thickness of the lining and the roughness of the surface in contact with the air.

5.0

REFERENCES

5.1

ASHRAE Handbook, Fundamentals, 2001 Edition

5.2

Handbook of Air-Conditioning System Design, Carrier Air-Conditioning Company, 1965 edition.

5.3

Advanced Design of Ventilation Systems for Contaminant Control by Howard D. Goodfellow, Elsevier Science Publishers, B. V. 1985.

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DESIGN GUIDE FOR AIR WASHERS

SHEET i OF iii

DESIGN GUIDE FOR AIR WASHERS

FILE NAMES: M6ME311R2.DOC AND M6ME311R2.DWG

REV. NO.

R0

R1

R2 ISSUE

INITIALS

SIGN.

INITIALS

SIGN.

INITIALS

PPD. BY

SM/GSJ

Sd/-/Sd/-

IYM/HRK

Sd/-/Sd/-

IYM

CHD. BY

SJB/SCM

Sd/-/Sd/-

PRJ

Sd/-

TPR

APD. BY

DHNR

Sd/-

RL

Sd/-

RL

SIGN.

INITIALS

SIGN.

R2

DATE

19.12.1980

24.03.2000

20.03.2003 TCE FORM NO. 020R2

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SECTION: CONTENTS

TCE.M6-ME-811-311

DESIGN GUIDE FOR AIR WASHERS

SHEET ii

OF iii

CONTENTS SL. NO.

TITLE

SH. NO.

1.0

SCOPE

1

2.0

APPLICATIONS

1

3.0

TYPES OF AIR WASHERS

7

4.0

PROCEDURE FOR SIZING

9

5.0

REFERENCES

11

1.

RECIRCULATED WATER SPRAY OR EVAPORATIVE AIR COOLING

2

2.

PRE-HEATED AIR WITH RECIRCULATED WATER SPRAY

3

3.

HEATED WATER SPRAY

4

4.

COOLING AND DEHUMIDIFICATION

6

5.

TYPICAL SPRAY TYPE AIR WASHER

12

6

TYPICAL RIGID MEDIA PAD TYPE AIR WASHER

12

FIGURES

TABLE 1. SATURATION EFFICIENCIES AND LENGTHS OF AIR WASHERS

8

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TCE.M6-ME-811-311

REVISION STATUS REV. NO.

DATE

DESCRIPTION

R0

19.12.1980

----------------

R1

24.03.2000

Generally revised.

R2

20.03.2003

Generally revised.

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1.0

SHEET 1 OF 12

SCOPE Air washer is a chamber or enclosed casing where water and air come in direct contact for applications like humidification, cooling & dehumidification and air cleaning. This design guide briefly describes these applications, describes various types of air washers available and gives procedure for sizing of most commonly used air washers.

2.0

APPLICATIONS

2.1

HUMIDIFICATION Humidification process increases the specific humidity of the air. Figures 1, 2 and 3 illustrate thermodynamic changes that occur between the air and water that are in direct contact. Humidification can be achieved in three ways :

2.1.1

(a)

Using recirculated water spray

(b)

Pre-heating the air and

(c)

Using heated water spray

Recirculated Water Spray Recirculated water spray (also called evaporative air cooling) reduces the air Dry Bulb Temperature (DBT) by evaporation of water into an air stream. Figure 1 illustrates thermodynamic changes that occur between the air and water that are in direct contact. The continuously recirculated water reaches an equilibrium temperature equal to the Wet Bulb Temperature (WBT) of the entering air. The heat and mass transfer between the air and water lowers the air DBT and increases the specific humidity at a constant WBT. Evaporative air cooling or saturation efficiency (Ec) is given by the relation: =

T1 – T2 100 x T1 – T3

Ec

=

Saturation efficiency

T1

=

DBT of the entering air

T2

=

DBT of the leaving air

T3

=

WBT of entering air

Ec Where,

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Evaporative air cooling system is normally provided in areas where wet bulb depression is more than 11OC. In cases where inside heat gain is very high, irrespective of the wet bulb depression, evaporative cooling is provided as this will reduce the required air flow rate. Figure 1 represents this schematically on a psychrometric diagram. Figure 1 also indicates the thermodynamic process showing the air temperature variation and constant water temperature. AIR WASHER AIR OUTLET CONDITION

WET BULB TEMPERATURE AIR WASHER AIR INLET CONDITION

SPECIFIC HUMIDITY (gm / KG OF DRY AIR)

SATURATION CURVE

T3

T2

T1

DRT BULB TEMPERATURE (0 C)

FINAL AIR TEMPERATURE DIFFERENCE (T2 – T3)

INITIAL AIR TEMPERATURE DIFFERENCE (T1 – T3)

INLET AIR TEMPERATURE T1 OUTLET AIR TEMPERATURE T2

WATER TEMPERATURE T3 (CONSTANT)

FIGURE 1 RECIRCULATED WATER SPRAY OR EVAPORATIVE AIR COOLING

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2.1.2

SHEET 3 OF 12

Pre-heating the Air Air is pre-heated before it enters the air washer. Pre-heating the air increases both DBT and WBT and lowers Relative Humidity (RH), but does not alter the specific humidity. At a higher WBT and with the same specific humidity, more water can be absorbed per unit mass of dry air passing through the air washer. Process that occurs in the air washer is the same as that for recirculated water spray. The final preferred conditions are achieved by adjusting the amount of pre-heating to give the required WBT at the entrance of air washer. Figure 2 represents this process schematically on a psychrometric diagram. AIR WASHER AIR OUTLET CONDITION WET BULB TEMPERATURE AIR WASHER AIR INLET CONDITION EVAPORATIVE COOLING SPECIFIC HUMIDITY (gm / KG OF DRY AIR)

SATURATION CURVE

HEATER ENTRY AIR CONDITION

T0

T3 T2

T1

AIR HEATING

DRT BULB TEMPERATURE (0 C)

FIGURE 2 PRE-HEATED AIR WITH RECIRCULATED WATER SPRAY

2.1.3

Heated Water Spray Spray water is heated before it is sprayed in the air washer. It is possible to raise both DBT and WBT above the DBT of the entering air. The leaving air condition may be controlled by (a)

Bypassing some of the air around the air washer and re-mixing the two air streams downstream or

(b)

Automatically reducing the number of operating spray nozzles.

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Figure 3 represents this process schematically on a psychrometric diagram. Also indicated is the relative variation of the air and water temperature in the process.

SATURATION CURVE AIR WASHER AIR OUTLET CONDITION

SPECIFIC HUMIDITY (gm / KG OF DRY AIR)

WET BULB TEMPERATURE AIR WASHER AIR INLET CONDITION

T3

T1

T2

DRT BULB TEMPERATURE (0 C)

FINAL TEMPERATURE DIFFERENCE (T4 – T2)

INITIAL AIR TEMPERATURE DIFFERENCE (T3 – T1)

INLET WATER TEMPERATURE T3 OUTLET WATER TEMPERATURE T4

CONTACT TIME

OUTLET AIR TEMPERATURE T2 INLET AIR TEMPERATURE T1

FIGURE 3 HEATED WATER SPRAY

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2.2

SHEET 5 OF 12

COOLING AND DEHUMIDIFICATION Cooling and Dehumidification reduces the DBT and specific humidity. Figure 4 represents this process schematically on a psychrometric diagram. Also indicated is the relative variation of the air and water temperature in the process. Heat and moisture removed from the entering air raises the water leaving temperature. Dehumidification results if the leaving water temperature is below the entering dew-point temperature. Moreover, the final water temperature is determined by the sensible and latent heat pick-up and the quantity of water circulated. Air leaving the air washer is substantially saturated. The difference between DBT and WBT is less than 0.5°C. The difference between leaving air and leaving water temperatures depends on the difference between entering DBT and WBT and the construction features of the air washer. The rise in water temperature is usually between 3 to 7OC. The most common air washer arrangement for cooling and dehumidifying air has two spray banks and is 2,500 to 3,500 mm long. If the air washer can cool and dehumidify the entering air to a WBT equal to the leaving water temperature, the performance factor of such an air washer is said to be 1.0. The actual performance of any air washer is the actual enthalpy change divided by the enthalpy change in washer having a performance factor of 1.0. The required performance factor may be calculated by: h1 – h2 Fp = h1 – h3 Where, Fp

=

Performance factor

h1

=

Enthalpy at entering air WBT

h2

=

Enthalpy at leaving air WBT at actual condition

h3

=

Enthalpy at WBT leaving the air washer with Fp = 1.0 i.e. at WBT equal to leaving water temperature

This type of system was used extensively for applications having large air flow capacities, e.g. synthetic fibre plants. This is an open system and ingress of dust and particulate contaminants from the process, which may be absorbed by the water, may cause clogging of chiller tubes and also the piping system. This could lead to major maintenance problems, breakdown and loss of production. Hence this type of system is normally not preferred.

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SPECIFIC HUMIDITY (gm / KG OF DRY AIR)

WET BULB TEMPERATURE

SATURATION CURVE

AIR WASHER AIR OUTLET CONDITION

AIR WASHER AIR INLET CONDITION

T2

T3

T1

DRT BULB TEMPERATURE (0 C)

FINAL TEMPERATURE DIFFERENCE (T2 – T4)

INITIAL TEMPERATURE DIFFERENCE (T1 – T3)

INLET AIR TEMPERATURE T1 OUTLET AIR TEMPERATURE T2

CONTACT TIME

INLET WATER TEMPERATURE T3

OUTLET WATER TEMPERATURE T4

FIGURE 4 COOLING AND DEHUMIDIFICATION

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2.3

SHEET 7 OF 12

AIR CLEANING Air washers are also used for removing dust particles etc. before it enters the process or discharged to atmosphere. The dust removal efficiency of air washer depends largely on the size, density, wettability and solubility of the dust particle. Efficiency is higher for the larger and the more wettable particles, separation being accomplished entirely by impingement on the wetted surfaces of the eliminator plates. Experience has shown that the spray itself is relatively ineffective in removing most atmospheric dusts. Air washers are of little use in removing soot particles because of the absence of an adhesive effect from the greasy surface. They are also ineffective in removing smoke because of inadequate inertia of the small particles (less than 1 micron). Rigid media pad type washers, however are efficient air cleaners. In practice they remove from 70 to 90 % percent by weight of air borne solid particles which include most of the particles exceeding 5 microns and many down to size 1 micron. The actual size and arrangement should be verified based on manufacturer’s data.

3.0

TYPES OF AIR WASHERS

3.1

SPRAY TYPE AIR WASHERS Spray type air washer consists of an air-tight chamber or casing containing air distribution plates, spray nozzle system arranged in a single bank or multiple banks, a tank for collecting the spray water, eliminators for removing entrained drops of water from the air, flooding nozzles for cleaning eliminators and a pump with piping, valves and specialities for recirculating water at higher rate than the evaporation rate. The tank is provided with drain, overflow, make-up water and quick-fill connections. Drain sump with suction screens are provided in tank. Spray type air washers are usually available from 3,500 to 4,25,000 M3 /Hr capacity. However, there is practically no limit for specially constructed air washers. There is no standardisation in air washer sizes. Normally, recirculated water flow rate is 0.53 M3 /Hr for a single bank and 1.34 M3 /Hr for double or multiple banks per 1,000 M3 /Hr of air flow rate. Air resistance varies from 6 mmWC to as high as 25 mmWC. Spray nozzle pressure varies from 1.4 to 2.8 Kg/cm2 g, whereas flooding nozzle operates at 0.2 to 0.3 Kg/cm2 g pressure. Generally, specified air face velocity through the air washer internals i.e. distribution plate, spray nozzles and the eliminators is 2.5 M/sec. Typical spray type air washer is shown in Figure 5.

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Tank has a minimum depth of 600 mm with 450 mm water level. The distance between air distribution plates and spray bank as well as spray bank and eliminator plates is 400 to 600 mm. The distance between the two spray banks is 800 to 1,400 mm. Normally, air washer tank is extended beyond eliminators by 500 mm. An air-tight inspection door is provided on the chamber or casing along with a marine light and catwalk inside the spray section, to facilitate cleaning of spray and flooding nozzles during maintenance. It is recommended that a level switch be provided to safeguard the pump. However, this is left to the discretion of the designer as the air washer operation is normally manual and the operator verifies and monitors the adequacy of the water level in the basin to ensure proper pump operation. For details and selection of air washers for humidification or cooling and dehumidification applications, refer Handbook of Air-conditioning System Design, Carrier Air-Conditioning Company, 1965 edition, Part 6, Chapter 2. Air washers are available in metallic or Fibre Reinforced Plastic (FRP) construction for smaller capacities upto about 25,000 M3 /Hr. For higher capacity these are normally in masonary construction. The saturation efficiencies and lengths of air washers are given in table 1 below: TABLE 1 SATURATION EFFICIENCIES AND LENGTHS OF AIR WASHERS SL. NO.

SPRAY BANK ARRANGEMENT

SATURATION EFFICIENCY %

LENGTH OF AIR WASHER M

1.

Single bank facing downstream side of air washer

50 to 60

1.2

2.

Single bank facing upstream side of air washer

65 to 80

1.8

3.

Double bank both facing downstream side of air washer

80 to 90

2.4 to 3

4.

Double bank with one facing downstream side and other facing upstream side of air washer

85 to 95

2.4 to 3

5.

Double bank both facing upstream side of air washer

90 to 98

2.4 to 3

6.

Multiple bank

90 to 98

3.5

SOURCE : ASHRAE HANDBOOK, HVAC SYSTEMS AND EQUIPMENT, 2000 EDITION, CHAPTER 19, PAGE 19.7

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3.2

SHEET 9 OF 12

HIGH VELOCITY SPRAY TYPE AIR WASHER High velocity spray type air washers generally operate at air face velocities in the range of 6 to 9 M/sec. However, velocity as high as 12 M/sec may also be used. The reduced cross-sectional area allows air washers to be installed in a less space. High velocity spray type air washers are rectangular in cross-section and except for eliminators are similar in appearance and construction to spray type air washers. Eliminator blades come in varying shapes, but most are a series of aerodynamically clean, sinusoidal shapes. Air resistance varies from 15 mmWC to as high as 40 mmWC.

3.3

RIGID MEDIA PAD TYPE AIR WASHER Rigid media pad type air washer consists of an air-tight chamber or casing of Galvanised Steel Sheet (GSS) or FRP containing corrugated rigid media pad, a tank for collecting water and a pump with piping, valves and specialities for recirculating water.Rigid media pad shall be of cellulose and fibre glass that have been treated chemically with antirot and rigidifying resins. Details for these media and fill configurations may be referred from vendor catalogues. Pad shall be cross-corrugated to maximise mixing of air and water in the direction of air flow. Depth of pad shall be 300 to 600 mm. The pad shall be arranged in tiers over which water is distributed. No atomisation of water is required. Only good distribution of water over the face of the pad is essential. Nozzles operate under comparatively low pressure as compared to those in spray type air washer. It is recommended that fine filters, with a filtering efficiency of 99% down to a particle size of 5 microns with test dust 2 as per IS 7613, in conjunction with pre-filters, with a filtering efficiency of 90% with test dust 3 as per IS 7613, be provided with rigid media type air washers. This is to minimise the clogging of the media with dust carried by the incoming air flow. However, this aspect may be examined by the designer with the site conditions in the respective applications. Saturation efficiency of this type of air washer is 70 to 95 %. Figure 6 shows a typical rigid media pad type air washer.

4.0

PROCEDURE FOR SIZING Procedure for evaporative air cooling application is given below :

4.1

INPUT DATA DBT of the entering air in OC

=

T1

ISSUE R2

TCE FORM NO. 120 R1

TCE CONSULTING ENGINEERS LIMITED

SECTION: WRITE-UP

TCE.M6-ME-811-311

DESIGN GUIDE FOR AIR WASHERS

4.2

WBT of the entering air in OC

=

T3

Heat load in KCal/Hr (calculated as per TCE.M6-ME-811-309)

=

Q

DBT to be maintained inside the space

=

Ta

SHEET 10 OF 12

LEAVING AIR TEMPERATURE Saturation efficiency Ec =

T1 – T2 T1 – T3

The value of saturation efficiency Ec may be referred to from para 3.1. Knowing the saturation efficiency, T2 can be calculated. Where, T2 is the DBT of the air leaving the air washer. 4.3

TEMPERATURE RISE DUE TO FAN HEAT GAIN The temperature rise due to fan heat gain is calculated by: P x 9.81 x 1.15 ∆T = η x 10 x ρa Where, P

=

η

=

Static pressure of fan in mmWC (calculated as per TCE.M1ME- 811-302) Fan static efficiency (normally assumed @ 60%)

ρa

=

Air density in Kg/M3

4.4

AIR FLOW RATE

4.4.1

For draw-through type air washer Air flow rate in M3 /Hr =

4.4.2

For blow-through type air washer Air flow rate in M3 /Hr =

4.5

Q 0.288 [ Ta - (T2 + ∆T) ]

Q 0.288 [ (Ta + ∆T) - T2 ]

RECIRCULATING WATER FLOW RATE In the absence of manufacturer’s data, para 3.1 above may be referred for calculating recirculated water flow rate. Knowing the re-circulating water flow rate, pumps capacity selection and piping sizing may carried out.

ISSUE R2

TCE FORM NO. 120 R1

TCE CONSULTING ENGINEERS LIMITED

SECTION: WRITE-UP

TCE.M6-ME-811-311

DESIGN GUIDE FOR AIR WASHERS

4.6

SHEET 11 OF 12

FACE AREA OF AIR WASHER Face area of the air washer in M2

=

Air flow rate in M3 /Hr 3600 x air face velocity (M/sec)

Depending on the available space, the height (H) and the width (W) of the air washer may be arranged such that face area A=WH. It is preferable to arrange the face area of the air washer such that width and height are equal. The length of the double bank air washer may be minimum 2.5 M and may be as long as 3.5 M. 4.7

MAKE-UP WATER FLOW RATE Make-up water flow rate varies from 1 to 1.5 % of recirculating water flow rate. However, this may be specifically calculated considering evaporation and blowdown. Blowdown rate shall be calculated considering the upper limit to be maintained in the concentration of dissolved solids in the recirculated water. Another criteria for the blowdown water quantity, are the scaling characteristics of the recirculated water. Generally, the scaling criterion is not applicable for air washers, as there is no change in the operating temperature of water for recirculated spray air washers. However, this criterion may be significant for pre-heated air and heated water spray air washers. The blowdown quantity and the scaling criteria estimated from the Puckorious Scaling Index (PSI) shall be calculated from the methods given in para 5.4.1 and para 5.3 respectively of TCE.M1-ME-127-201, Basic Study Guide for Cooling Tower Make-up System.

5.0

REFERENCES

5.1

ASHRAE Handbook, HVAC Systems and Equipment - 2000 Edition

5.2

Handbook of Air-conditioning System Design, Carrier Air-Conditioning Company, 1965 edition

5.3

TCE.M1-ME-127-201 Basic Study Guide for Cooling Tower Make-up System

5.4

TCE.M1-ME-811-302 Basic Study Guide for Ventilation System

5.5

TCE.M6-ME-811-301 Design Guide for Air-conditioning Load Calculations

5.6

TCE.M6-ME-811-309 Design Guide for Ventilation Heat Load Calculations

ISSUE R2

TCE FORM NO. 120 R1

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR REFRIGERATION LOAD CALCULATIONS FOR PROCESS COOLING

TCE.M6-ME-811-313

SECTION: TITLE SHEET i OF iii

DESIGN GUIDE FOR REFRIGERATION LOAD CALCULATIONS FOR PROCESS COOLING

FILE NAME: M6ME313R3.DOC REV. NO.

R0

R1

R2

R3 ISSUE

INITIALS

SIGN.

INITIALS

SIGN.

INITIALS

SIGN.

INITIALS

PPD. BY

SCM

Sd/-

TPR

Sd/-

IYM

Sd/-

RRC

CHD. BY

SJB

Sd/-

PRJ

Sd/-

PRJ

Sd/-

TPR

APD. BY

DHNR

Sd/-

RL

Sd/-

RL

Sd/-

RL

SIGN.

R3

DATE

19.12.1980

10.12.1997

02.03.2001

01.03.2004 TCE FORM NO. 020R2

TCE.M6-ME-811-313

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR REFRIGERATION LOAD CALCULATIONS FOR PROCESS COOLING

SECTION: CONTENTS SHEET ii OF iii

CONTENTS SL. NO.

TITLE

SH. NO.

1.0

SCOPE

1

2.0

GENERAL

1

3.0

PROCESS COOLING LOAD

2

4.0

HEAT GAIN DUE TO SUPPLY AND RETURN COOLANT PIPING

2

5.0

HEAT GAIN DUE TO COOLANT PUMPS AND AGITATORS

4

6.0

HEAT GAIN IN MIXING TANK

5

7.0

REFRIGERATION LOAD

5

8.0

REFERENCES

6

ISSUE R3 TCE FORM NO. 120 R1

TCE.M6-ME-811-313

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR REFRIGERATION LOAD CALCULATIONS FOR PROCESS COOLING

SECTION: REV. STATUS SHEET iii OF iii

REVISION STATUS REV. NO.

DATE

DESCRIPTION

R0

19.12.1980

--

R1

10.12.1997

Document number changed. British units deleted and SI units incorporated. Generally revised.

R2

02.03.2001

Generally revised.

R3

01.03.2004

Para 8.0 revised and document reformatted.

ISSUE R3 TCE FORM NO. 120 R1

TCE.M6-ME-811-313

1.0

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR REFRIGERATION LOAD CALCULATIONS FOR PROCESS COOLING

SECTION: WRITE-UP SHEET 1 OF 6

SCOPE This design guide describes the method for calculating the refrigeration load on the refrigeration equipment for process cooling where a secondary coolant like chilled water or brine is used to remove heat from the various process equipment in industrial and chemical plants.

2.0

GENERAL The total load on the refrigeration equipment is the actual process cooling load plus the heat gain by circulating coolant pumps, supply and return coolant piping, mixing/storage tank, agitators etc. The following data shall be obtained from the client or process engineer or process equipment designer before proceeding with the calculations: (a)

Type of coolant i.e. chilled water or brine such as calcium chloride, sodium chloride, glycol etc. If client or process engineer or process equipment designer does not specifically indicate requirement of type of coolant, it may be recommended by refrigeration system designer and consent obtained from client or process engineer or process equipment designer.

(b)

Flow rate of coolant through various process equipment

(c)

Coolant supply and return temperatures with tolerances at battery limit or process equipment

(d)

Coolant supply and return pressures with tolerances at battery limit or process equipment

(e)

Locations of process equipment and refrigeration equipment (Refer plot plan and GA drawings)

(f)

Routing of supply and return coolant piping from refrigeration equipment to battery limit or process equipment (Refer piping layout drawings. In case piping layout drawings are not available, the refrigeration system designer shall assume the most probable pipe routing)

(g)

Diversity

(h)

Special requirements, if any

ISSUE R3 TCE FORM NO. 120 R1

TCE.M6-ME-811-313

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR REFRIGERATION LOAD CALCULATIONS FOR PROCESS COOLING

3.0

PROCESS COOLING LOAD

3.1

IN MKS UNITS

SECTION: WRITE-UP SHEET 2 OF 6

Process cooling load in Kcal/Hr = ρ × V × Cp × ∆t Where,

ρ

= Density, Kg/M3

V = Flow rate, M3 /Hr Cp = Specific heat, KCal/Kg OC ∆t = Temperature rise, OC 3.2

IN SI UNITS Process cooling load in KW = ρ × V × Cp × ∆t Where,

ρ

= Density, Kg/M3

V = Flow rate, M3 /Sec Cp = Specific heat, KJ/Kg OC ∆t = Temperature rise, OC 3.3

For chilled water, value of ρ is 1000 in both MKS and SI units and values of Cp are 1.0 and 4.1868 in MKS and SI units respectively. However, for brines the values of ρ and Cp depend upon the type, temperature and concentration of brines. The values of ρ and Cp for selected brines at various temperatures and concentrations may be obtained from the references cited in para 8.0. These properties have to be taken at mean operating temperature.

4.0

HEAT GAIN DUE TO SUPPLY AND RETURN COOLANT PIPING

4.1

IN MKS UNITS Heat transferred per unit area in KCal/M2 Hr, q =

ta - ti R

Total heat transferred in KCal/Hr, Q = q × AL Where,

ta = Ambient air temperature, OC ti =

Supply temperature of coolant, OC

ISSUE R3 TCE FORM NO. 120 R1

TCE.M6-ME-811-313

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR REFRIGERATION LOAD CALCULATIONS FOR PROCESS COOLING

SECTION: WRITE-UP SHEET 3 OF 6

R = Overall thermal resistance of the pipe with insulation, M2 HrOC/Kcal AL = Pipe outside area with insulation, π × DL × L, M2 DL = Outside diameter of pipe with insulation, M L = 4.2

Length of pipe, M

IN SI UNITS Heat transferred per unit area in W/M2 , q =

ta – ti R

Total heat transferred in KW, Q = (q × AL)/1000 Where,

ta = Ambient air temperature, OC ti =

Supply temperature of coolant, OC

R = Overall thermal resistance of the pipe with insulation, M2OK/W AL = Pipe outside area with insulation, π × DL × L, M2 DL = Outside diameter of pipe with insulation, M L =

Length of pipe, M

Overall thermal resistance of pipe, R = RI + RP + RL + RO R

=

DL hI DI

+

DL In (DO/DI) 2 kP

+

DL In (DL/DO) 2 kL

1 hO

+

Where the abbreviations used are as follows: ITEM

COOLANT

PIPE

PIPE

INSULATION

AIR

Diameter, M

--

Inside DI

Outside DO

Outside DL

--

Thermal conductivity, Kcal/MHrOC (W/MOK)

--

kP

kP

kL

--

Heat transfer coefficient, Kcal/M2 HrOC (W/M2OK)

hI

--

--

--

hO

ISSUE R3 TCE FORM NO. 120 R1

TCE.M6-ME-811-313

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR REFRIGERATION LOAD CALCULATIONS FOR PROCESS COOLING ITEM

SECTION: WRITE-UP SHEET 4 OF 6

COOLANT

PIPE

PIPE

INSULATION

AIR

RI

RP

RP

RL

RO

Resistance, M2 HrOC/KCal (M2OK/W)

The resistance of finishing cement layer or any other protective layer for insulation is neglected. Resistance of inside film, pipe material and outside film may also be neglected. These neglected components do not significantly affect the heat gain. For both supply and return piping, tc may be taken equal to the temperature required at the inlet of battery limit/process equipment. Thus R =

and q =

DL In (DL/DO) 2 kL

DL 2 kL

ta - tc ln (DL/DO)

4.3

Increase the value of ‘Q’ as calculated by 10% to account for heat gain through pipe fittings, valves etc.

5.0

HEAT GAIN DUE TO COOLANT PUMPS AND AGITATORS

5.1

The power required to pump the coolant or drive the agitators adds heat to the system. This may be estimated as follows: Heat added = BKW × 860 KCal/Hr (MKS units) = BKW (SI units) Where BKW is the power required to drive the pump/agitator.

5.2

In case the actual power to drive the pump is not available same may be calculated as given below: BKW = (V × H × SG)/(ηP × 367) Where,

V = Flow rate, M3 /Hr H = Total pump head, M SG = Specific gravity ηP = Pump efficiency

ISSUE R3 TCE FORM NO. 120 R1

TCE.M6-ME-811-313

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR REFRIGERATION LOAD CALCULATIONS FOR PROCESS COOLING

SECTION: WRITE-UP SHEET 5 OF 6

For small pumps of capacities 4.5 to 22 M3 /Hr the pump efficiency may be assumed as 50% and for pumps greater than 22 M3 /Hr it may be assumed as 70%. 5.3

In the absence of BKW of agitator, heat equivalent of 80% of driver motor rating for agitator may be taken as heat added to coolant.

6.0

HEAT GAIN IN MIXING TANK

6.1

Surface of the tank through which there is heat transfer shall be considered and the heat gain shall be calculated by the following formula: q = (ta - tc)/(Z/kL) Q = q × AL, Kcal/Hr (MKS units) = (q × AL)/1000, KW (SI units) Where,

Z = Insulation thickness, M AL = Total outside surface area with insulation, M2

6.2

For other symbols and assumptions, refer para 4.0.

7.0

REFRIGERATION LOAD

7.1

The cooling load on the refrigeration equipment is the sum of the process cooling load (para 3.0) and the heat gain due to piping, pumps and agitators, and mixing tank (paras 4.0 + 5.0 + 6.0). The diversity factor and special requirements, if any, shall also be considered.

7.2

Apart from the refrigeration load, the temperatures of coolant required at the battery limit/process equipment are also important. The heat gain due to pumps, piping, tank etc. raises the temperature of the coolant. Therefore, the temperature at the outlet of the chilling unit shall be lower than that required at the inlet of the battery limit or process equipment. The temperatures at the outlet and inlet of the chilling unit are calculated or specified as indicated in para 7.3.

7.3

The heat gain in the return coolant piping is a small percentage of the total heat gain due to pumps, tank and supply piping. Therefore, for calculating the temperature at the outlet of the chilling unit the entire heat gain is considered to be on the supply side.

7.4

IN MKS UNITS Temperature rise in OC, ∆t =

QT

ρ × V × Cp ISSUE R3 TCE FORM NO. 120 R1

TCE.M6-ME-811-313

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR REFRIGERATION LOAD CALCULATIONS FOR PROCESS COOLING

Where,

SECTION: WRITE-UP SHEET 6 OF 6

QT = Heat gain due to piping, pumps, tanks etc. (paras 4.0 + 5.0 + 6.0), KCal/Hr ρ

= Density, Kg/M3

V = Flow rate, M3 /Hr Cp = Specific heat, KCal/Kg OC 7.5

IN SI UNITS Temperature rise in OC, ∆t = Where,

QT ρ × V × Cp

QT = Heat gain due to piping, pumps, tanks etc. (paras 4.0 + 5.0 + 6.0), KW ρ

= Density, Kg/M3

V = Flow rate, M3 /Sec Cp = Specific heat, KJ/Kg OC The temperature to be specified at the outlet of the chilling unit shall be lower by ∆t than that at the inlet to battery limit or process equipment. The temperature to be specified at the inlet of the chilling unit shall be the same as that at the outlets of battery limit or process equipment. 8.0

REFERENCES

8.1

2001 ASHRAE Handbook - Fundamentals

8.2

Handbook of Air Conditioning System Design - Carrier Air Conditioning Company - 1965 edition.

ISSUE R3 TCE FORM NO. 120 R1

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR HEAT RECOVERY WHEEL

TCE.M6-ME-811-314

SECTION: TITLE

SHEET i OF iii

DESIGN GUIDE FOR HEAT RECOVERY WHEEL

FILE NAME: M6ME314R0.DOC AND M6ME314R0.DWG FLOPPY NO :TCE-971-ME-FP-101

REV.NO.

R0 INITIALS

PPD.BY

SS

CHD.BY

ST

APD.BY

RNW/RL

DATE

ISSUE SIGN.

INITIALS

SIGN.

INITIALS

SIGN.

INITIALS

SIGN.

R0

01.08.16 FORM NO. 020R2

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR HEAT RECOVERY WHEEL

TCE.M6-ME-811-314

SECTION:CONTENTS

SHEET

ii OF iii

CONTENTS

SL.NO.

TITLE

SH.NO.

1.0

SCOPE

1

2.0

AIR TO AIR ENERGY RECOVERY DEVICES

1

3.0

HEAT RECOVERY WHEEL - OPERATING PRINCIPLE

1

4.0

CONSTRUCTION

2

5.0

PERFORMANCE RATING

7

6.0

PSYCHROMETRIC ANALYSIS

10

7.0

APPLICATIONS

10

8.0

REFERENCES

11

APPENDICES 1.

APPENDIX-1 -

COMPARISON OF AIR TO ENERGY RECOVERY DEVICES.

AIR

13

2.

APPENDIX-2 - SELECTION OF HEAT RECOVERY WHEEL (WITH TABLE-1)

15

TABLE-1:

TYPICAL MANUFACTURER’S PERFORMANCE DATA FOR HEAT RECOVERY WHEEL (HRW)

22

3.

APPENDIX-3 - CONFIGURATION OF HRW & PURGE SECTION

23

4.

APPENDIX-4 -

24

5.

APPENDIX-5 - PSYCHROMETRIC PROCESS OF HEAT RECOVERY USING HEAT WHEEL

25

6.

APPENDIX-6 -

26

INSTALLATION GUIDE LIN ES (HRW)

TYPICAL HRW CATALOGUE DATA PHYSICAL DIMENSIONS OF A HRW

ISSUE R0 FORM NO. 120 R1

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SECTION:REV.STATUS

SHEET

iii OF iii

REVISION STATUS

ISSUE R0 FORM NO. 120 R1

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR HEAT RECOVERY WHEEL

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1.0

SECTION:WRITE UP

SHEET

1 OF 26

SCOPE This document describes design and construction features of Heat Recovery Wheels for air conditioning application and provides guidelines for their selection.

2.0

TYPES OF AIR TO AIR ENERGY RECOVERY DEVICES

2.1

ASHRAE Equipment Handbook 1996 refers to the following six types of air-to-air heat exchange devices. ♦ ♦ ♦ ♦ ♦ ♦

Rotary Energy Exchangers (Enthalpy or Heat Recovery Wheel) Coil Energy Recovery Loop (Run around coil loop) Twin-Tower Enthalpy Recovery Loop Heat Pipe Heat Exchangers Fixed Plate Exchangers Thermosyphon Heat Exchangers

2.2

Distinction is made between the sensible only and the total heat exchangers – total heat meaning sensible plus latent heat or enthalpy. The twin-tower loop is a total heat exchanger. The rotary exchanger, or heat recovery wheel, may be constructed as either a sensible only or a total heat device. The rest are essentially sensible heat exchangers in which transfer of latent heat, if any, is incidental.

2.3

Because of their ability to transfer both sensible and latent heat, the enthalpy devices are far more effective in energy recovery. It is found that under summer design conditions, the total heat device typically recovers nearly three times as much energy as the sensible heat device. Under winter conditions, it recovers over 25% more (ASHRAE: Equipment Handbook 1988).

2.4

From the comparison of Air-to-Air Energy Recovery Devices APPENDIX-1, it is seen that the enthalpy wheel has the highest effectiveness and the least pressure drop at any face velocity. The other enthalpy device – twin-tower loop – has the highest pressure drop for the same effectiveness.

3.0

HEAT RECOVERY WHEEL OPERATING PRINCIPLE

3.1

A rotary air-to-air energy exchanger, or heat recovery wheel, has a revolving cylinder filled with an air-permeable medium having a large inter surface area. Adjacent supply and exhaust airstreams each flow through one-half of the exchanger in a counterflow pattern (APPENDIX-3). Heat transfer media may be selected to recover only sensible heat or total heat. ISSUE R0 FORM NO. 120 R1

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3.2

Sensible heat is transferred as the medium picks up and stores heat from the hot airstream and releases it to the cold one. Latent heat is transferred as the medium (1) condenses moisture from the airstream with the higher humidity ratio (either because the medium temperature is below its dew point) or by means of adsorption for solid desiccants with a simultaneous release of heat; and (2) releases the moisture through evaporation (and heat pickup) into the airstream with the lower humidity ratio. The dessicant coated wheel slowly rotates between the outdoor and the return air stream. The higher temperature air stream gives sensible heat to wheel media. The energy is then, given up to the cooler air stream during the second half of the revolution. The moisture is captured and released by the dessicant coating which has an enormous internal surface area and strong attraction to water vapour. Since the opposing air streams have different temperatures and moisture contents, their vapour pressures differ. This vapour pressure differential serves as the driving force which causes the transfer of water vapour (latent energy). Thus, the humid air is dried while the dried air is humidified. In total heat transfer, both sensible and latent heat transfer occur simultaneously. Because heat recovery wheels have a counterflow configuration and normally use smalldiameter flow passages, they are quite compact and can achieve high transfer effectiveness.

4.0

CONSTRUCTION

4.1

Air contaminants, dew point, exhaust air temperature, and supply air properties influence the choice of structural materials for the casing, rotor structure, and medium of a rotary energy exchanger. Aluminium and steel are the usual structural, casing, and rotor materials for normal comfort ventilating systems. Exchanger media are fabricated from metal or mineral and are classified as providing either random flow or directionally oriented flow through their structures.

4.2

Random flow media are made by knitting wire into an open woven cloth or corrugated mesh, which is layered to the desired configuration. Aluminium mesh, commonly used for comfort ventilation systems, is packed in pie-shaped wheel segments. Stainless steel and monel mesh are used for high-temperature and corrosive applications. These media should only be used with clean, filtered airstreams because they plug easily. Random flow media also require a significantly larger face area than directionally oriented media for given values of airflow and pressure drop.

4.3

Directionally oriented media are available in various geometric configurations. The most common consist of small 1.5 to 2 mm (0.06 to 0.08 in.) air passages parallel to the direction of airflow. Whether the manufacturer uses triangular, hexagonal, or other shaped passages, ISSUE they are very similar in performance for a given flow passage. R0 FORM NO. 120 R1

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SECTION:WRITE UP

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Aluminium foil, inorganic sheet, treated organic sheet, and synthetic materials are used for low and medium temperatures. Stainless steel and ceramics are used for high temperatures and corrosive atmospheres. 4.4

Media surface areas exposed to airflow vary from 330 sq.m per cubic meter to 3300 sq.m per cubic meter (100 to over 1000 sq.ft/cu.ft), depending on the type of medium and physical configurations. Media may also be classified according to their ability to recover only sensible or heat or total heat. Media for sensible or total heat recovery are made of aluminium, copper, stainless steel, and monel and treated with a desiccant.

4.5

The enthalpy wheel is a cylinder, packed with a heat transfer medium that has numerous small air passages, or flutes, parallel to the direction of airflow. The flutes are triangular or semi-circular in cross-section. The structure, commonly referred to as the honeycomb matrix, is produced by interleaving flat and corrugated layers of a high conductivity material, usually aluminium, coated with a desiccant. Stainless steel, ceramic, and synthetic materials may be used, instead of aluminium, in specific applications. The flutes in most wheels measure between 1.5 mm to 2.0mm in height. Typical configuration of Heat Recovery Wheel is shown in APPENDIX-3.

4.6

In a typical installation, the wheel is positioned in a duct system such that it is divided into two half moon sections. Stale air from the conditioned space is exhausted through one half while outdoor air is drawn through the other half in a counter flow pattern. At the same time the wheel is rotated slowly (2 to 20 rpm). The rotor is supported by two pillow block bearings and is driven by an A.C. motor and permanently lubricated gear reducer. The rotor media is segmented for ease of field erection or replacement of one sector. The media is cleanable with low temperature steam, hot water or light detergent, without degrading the latent recovery. The recovery wheel casing is manufactured from tubular structure to provide a self supporting rigid structure complete with access panels, rotor, bearings, seals, drive mechanism with belt.

4.7

DESICCANT OPTIONS

4.7.1

The desiccant is a key element in the enthalpy wheel technology. It is relevant, therefore, to discuss briefly on the desiccant options currently available on wheels and their distinguishing features. A desiccant is a sorbent having particular affinity for water. Sorbents have been classified by ASHRAE (Fundamentals Handbook 1997, Chapter 19) into two basic categories, i.e. Absorbents and Adsorbents. The absorbents are substances that undergo a physical or a chemical ISSUE change in the process of attracting and holding moisture. An example R0 FORM NO. 120 R1

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SECTION:WRITE UP

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of this is table salt, which changes from a solid to a liquid as it absorbs moisture. Adsorbents, on the other hand, undergo no change except increase in weight to the extent of moisture adsorbed. However, absorbents are no more in use on enthalpy wheels. 4.7.2

Adsorbents are primarily solids with internal porosity of suc h order that the developed surface area of the pores is several hundred square meters per gram of the substance. Silica Gel, Activated Alumina, and Molecular Sieve – are the desiccants currently being offered on enthalpy wheels.

4.7.3

It is observed that molecular sieves have a relatively higher absorption capacity at low concentration levels of water vapors. However, the capacity does not increase significantly with increase in relative humidity. Both silica gel and activated alumina, on the other hand, have gradually increasing capacities, with silica gel adsorbing almost twice as much as molecular sieve at 100% RH. The decrease in adsorption capacity of molecular sieve with increase in temperature is much smaller compared to the other two desiccants.

4.7.4

These characteristics influence wheel design in terms of its desiccant mass, flute dimensions, rotational speed, etc., and determine moisture transfer effectiveness of the wheel at different temperature and humidity conditions of the two air streams. The overall performance of a model, however, depends on a number of other factors and must be established experimentally for each device. Manufacturers normally have detailed performance data on their equipment which must be consulted for a given application.

4.7.5

The adsorption behaviour of desiccants is regulated by (a) their total internal surface area, (b) the total volume of their capillaries, and (c) the range of their capillary diameters. A wide range of adsorption characteristics are possible within a single desiccant with variations in these parameters.

4.7.6

The desiccant is reactivated with building's exhaust air, and the air reactivates the dessicant adiabatically. No energy apart from what is contained in the exhaust air stream is required.

4.7.7

The desiccant of sufficient mass is coated with non masking porous binder adhesive so as to allow quick and easy uptake and release of water vapour. The weight of desiccant coating and the mass of aluminium foil (substrate) shall be in a ratio as to ensure equal recovery of both sensible and latent heat over the operating range.

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4.8

ADVANTAGES OF THE HEAT RECOVERY WHEEL

4.8.1

The Heat Recovery Wheel recovers both sensible and latent energy.

4.8.2

Latent recovery increases the energy savings potential as against sensible recovery only.

4.8.3

Ideal for applications where large quantities of fresh air are required.

4.8.4

Reduces outdoor air load by about 80%, (depending on inside and outside conditions) thus reducing first costs in capital equipment like chillers and boilers.

4.8.5

Unique desiccant surface selectivity eliminates adsorption of components other than water vapour.

4.8.6

Self-cleaning of rotor.

4.8.7

Better performance, higher efficiencies and lower pressure losses.

4.8.8

Reduced air volume requirements for purge and seal leakage.

4.8.9

Sectionalized media lends itself to a retrofit programme.

4.8.10

Special designs for sensible recoveries and for corrosive environments are available.

4.9

CROSS CONTAMINATION

4.9.1

Cross contamination or mixing of air between supply and exhaust airstreams occurs on all rotary energy exchangers by two mechanisms – carryover and leakage. Carryover occurs as air is entrained within the volume of the rotation medium and is carried into the other airstream, Leakage occurs because the differential static pressure across the two airstreams drives air from a higher to a lower static pressure region. Cross-contamination can be substantially reduced by placing the blowers so that they promote leakage of outside air to the exhaust airstream rather than the other way around. Carryover occurs each time a portion of the matrix passes the seals dividing the supply and exhaust airstreams. Because carryover from exhaust to supply may be undesirable, a purge section can be installed on the heat exchanger to reduce cross-contamination.

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4.9.2

In many applications, recirculating some air is not a concern. However, critical applications, such as hospital operating rooms, laboratories, and clean rooms, require stringent control of carryover. Carryover can be reduced by below 0.1% of the exhaust airflow with a purge section (ASHRAE SYSTEMS & EQUIPMENT HAND BOOK 1996).

4.9.3

The theoretical carryover of a wheel without a purge is directly proportional to the speed of the wheel and the void volume of the wheel media (75 to 95% void, depending on type and configuration).

4.9.4

For example, a 3 m diameter, 200 mm deep wheel with a 90% void and wheel volume operating at 14 rpm has a carryover volume of π( 3 / 2) 2 ( 0.2)( 0.9)(14) × 60 = 1070CMH ( approx.) If the wheel is handling a 34,000 CMH balanced flow, the percentage carryover is  1070   100 = 3.15%  34,000  The exhaust fan, which is usually located at the exit of the exchanger, should be sized to include leakage, purge and carryover airflows.

4.9.5

Heat recovery wheels are provided with a purge section as a standard design to limit trapped exhaust air from being carried over in the thermal transfer media to supply side. Using a purge section, a small stream of supply air is diverted, sent back through the wheel media and exhausted to the outside. By maintaining a pressure difference between the supply and exhaust ducts, exhaust carry over can be minimised. The schematic purge operation is shown in APPENDIX-3.

4.9.6

In order to compensate for the outside air being exhausted through the purge, the exhaust fan capacity should be increased depending on the pressure differential between the downstream supply side and the downstream exhaust side.

4.9.7

By utilising the correct fan arrangement, the seal and purge system can operate at optimum efficiency, virtually eliminating cross contamination. Installation guidelines are shown in APPENDIX-4.

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5.0

PERFORMANCE RATING

5.1

ASHRAE STANDARD 84, Method of Testing Air-to-Air Heat Exchangers, establishes rating and testing procedures for commercial air-to-air heat recovery devices.

5.2

ASHRAE STANDARD 84 defines effectiveness as E = Actual transfer (of moisture or energy) Max. possible transfer between air streams

…….. (1)

Where E = Sensible, latent or total heat effectiveness Refering to Fig 1, E = WS ( X 2 − X 1 ) / WMIN ( X 3 − X1 ) =

We ( X 3 − X 4 ) ……… (2) WMIN ( X 3 − X 1 )

where X = humidity ratio, dry bulb temperature or total enthalpy respectively at the locations 1,2, 3 & 4. Ws

= Mass Flow rate of Supply air, kg of dry air.

We

= Mass Flow rate of Exhaust Air, kg of dry air.

Wmin

= Minimum value Ws or We.

W  The leaving supply air condition is X 2 = X1 + E  MIN ( X 3 − X1 ) ...(3)  WS  And the leaving exhaust air condition is W  X 4 = X 3 − E  MIN ( X 3 − X1 )  …(4)  We  (SOURCE : ASHRAE 1996 SYSTEMS & EQUIPMENT HANDBOOK/ CHAPTER 42)

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X1 , WS 1. Supply Air Entering Energy Recovery device

We, X3 3. Exhaust Air Entering

FIG.1

X4 4. Exhaust Air Leaving

AIR STREAM NUMBERING CONVENTION

5.3

Air flow Arrangements and Effectiveness.

5.3.1

Heat Recovery Wheel effectiveness depends to a great extent on the air direction and pattern of the supply and exhaust airstreams. Parallel flow exchangers (Fig.2) have a theoritical max. effectiveness of 50%. Counterflow exchangers (Fig.3) can have an effectiveness approaching 100% but units designed for typical applications have a much lower effectiveness. The normal range is 50 to 85%. A median value of 80% can be assumed for design.

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Supply air

Energy Recovery Device Return Air

Exhaust Air

FIG 2 : PARALLEL HEAT EXCHANGE

Supply Air

Outside Air Energy Recovery device

Return Air

Exhaust Air

FIG .3 : COUNTER FLOW HEAT EXCHANGE

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PRESSURE DROP The pressure drop for each air stream through a heat recovery wheel depends on many factors, including heat exchanger design, mass flow rate, temperature, moisture and inlet/outlet air. The exchanger pressure drop must be overcome by fans. However the actual pressure drop shall be obtained from the manufacturer. Typical pressure drop value is indicated in APPENDIX-1.

6.0

PSYCHROMETRIC ANALYSIS In the skeleton Psychrometric Chart shown in APPENDIX-5, outdoor (fresh) air at 40.4°C DB/25.6°C WB (15 grams/kg of dry air) (Ab. Humidity) is drawn through the wheel (F1 ). Heat from return air from space at 24°C/ 9.5 grams /kg of dry air (50% RH) (F3 ) is recovered. Fresh air entering the space is cooled and dehumidified. Considering an effectiveness of 80%, supply condition F2 will be 27.3 °C /10.6 gms per kg. of dry air as per calculations shown below. 0.8 =

40.4 − X 2 40.4 − 24

.. . 40.4 – X2 = 0.8 (16.4) .. . X2 = 27.28°C Say 27.3 °C Similarly, 0.8 =

15 − X 2 15 − 9.5

or 15 – X2 = 4.4 gms/kg of dry air. X2 = 10.6 gms/kg of dry air. 7.0

APPLICATIONS

7.1

With the view to eliminate “Sick Building Syndrome” created due to reduced outdoor air (fresh air) quantities, ASHRAE STANDARD 62-1989 for Indoor Air Quality came out with increased ventilation rates. Reproduced below are the recommended ventilation rates under the ASHRAE 62-1989 standard for some of the applications : ISSUE R0 FORM NO. 120 R1

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RECOM. VENTILATION RATE/PERSON (L/S) 10 (20 CFM) 15 (30 CFM) per room 12.5 (25 CFM) 15 (30 CFM) 7.5 (15 CFM) 30 (60 CFM)

(Conversion : 1 CFM ~ 0.5 L/SEC.) 7.2

The most widespread application of enthalpy (Heat) wheels is for PRECONDITIONING fresh outside air before it is introduced to a building. Because the system is capable of recovering 80% of the heating or cooling energy that is exhausted from a building, the cost of fresh air ventilation is reduced. However, an economic analysis has to be made before recommending HRW for particular application.

7.3

Some typical applications where the Heat Wheel can be employed to reduce the fresh air load and improve the IAQ are : ♦ ♦ ♦ ♦ ♦ ♦ ♦

Hotels, Restaurants. Hospitals, Nursing Homes Conference rooms Airports Auditoriums Offices Health Clubs, etc.

7.4

A typical calculation procedure for selection of Heat Recovery Wheel for preconditioning outdoor air for air conditioning application, with a sample problem has been presented in APPENDIX-2.

8.0

REFERENCES

8.1

ASHRAE Hand Book – HVAC SYSTEMS & EQUIPMENT 1996.

8.2

ASHRAE STANDARD 62-1989.

8.3

ECO-Fresh – Technical Data, Arctic India. ISSUE R0 FORM NO. 120 R1

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8.4

“Carnes” Heat Wheel Catalogue.

8.5

ASHRAE JOURNAL OCT 1999 - "Evaluating Active Desiccant Systems for Ventilating Commercial Buildings" by L.G.Harriman & Others".

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APPENDIX-1 COMPARISON OF AIR-TO-AIR ENERGY RECOVERY DEVICES

Fixed Plate Heat Exchangers Counterflow Crossflow Parallel flow

Heat Recovery Wheel Counterflow Parallel flow

Heat Pipe Heat Exchangers Counterflow Parallel flow

Counter flow Parallel flow

Counterflow Parallel flow

Equipment Size range, CMH (cfm) Type of heat transfer (typical effectiveness)

85 (50) and up

85 to 119000 (50 to 70,000)

170 (100) and up

170 (100) and up

170 (100) and up

Sensible to 80%)

Sensible (50 to 80%) Total (55 to 85%)

Sensible to 65%)

Face velocity, m/min (fpm) (most common design velocity)

30.5 to 305 (61 to 305) 100 to 1000 (200 to 1000)

150 to 305 (500 to 1000)

120 to 240 (135 to 170) 400 to 800 (450 to 550)

Pressure drop, Pa of water (most likely pressure) (in. of water)

5Pa to 450 Pa (25Pa to 380 Pa) 0.02 to 1.8 (0.1 to 1.5)

100 to 180 (0.4 to 0.7)

Temperature range ºC (ºF)

-21 to 800 (-70 to 1500 °F)

Typical mode of purchase

1.Exchanger only 2. Exchanger with enclosure or casing 3. Exchanger and blowers Complete system

Air flow arran -gements

(50

Sensible to 65%)

(55

Thermosiphon

Twin Towers Enthalpy recovery loop

Sensible (40 to 60%)

Sensible (40 to 60%)

90 to 180 (300 to 600)

120 to 240 (135 to 170) 400 to 800 (450 to 550)

90 to 135 (300 to 450)

100 to 500 (0.4 to 2.0)

100 to 500 (0.4 to 2.0)

100 to 500 (0.4 to 2.0)

180 to 305 (0.7 to 1.2)

-21 to 800 (-70 to 1500 °F)

-4.4 to 35 (-40 to 95 °F)

-10 to 480 (-50 to 900 °F)

-4.4 to 40 (-40 to 104°F)

-4.4 to 46 (-40 to 115°F)

1

1. Exchanger only 2.Exchanger with enclosure or casing

1.Coil only 2.Complete system

1.

1.

Exchanger only 2. Exchanger with enclosure or casing 3. Exchanger and blowers Complete system

(45

Runaround Coil Loop

2.

Exchanger only Exchanger with enclosure or casing

Complete system

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APPENDIX-1 (CONTD..) Fixed Plate Heat Exchangers 1. No moving parts 2. Low pressure drop 3 Easily cleaned

Heat Recovery Wheel 1. Latent transfer 2. Compact large sizes 3. Low pressure drop

Heat Pipe Heat Exchangers 1. No moving parts except tilt . 2. Fan location not critical 3. Allowable pressure differential upto 60 in. of water

Limitations

Latent avail able in hygroscopic units only

Cross-leakage

0 to 5%

Cold climates may increase service Cross air contamination possible 1 to 10%

Heat rate control (HRC) schemes

Bypass dampers ducting

Wheel speed control over full range

Unique advantages

and

Runaround Coil Loop 1.

Thermosiphon

Twin Towers Enthalpy recovery loop 1. Latent transfer from remo te airstreams 2. Multiple units in a single system. 3. Efficient micro – biological cleaning of both supply and exhaust airstreams. Few Suppliers

Exhaust airstream can be separated from supply air Fan location not critical

1. No moving parts. 2. Exhaust airstream can be separated from supply air 3. Fan location not critical

Effectiveness limited by pressure drop and cost Few suppliers

High effective –ness requires accurate simulation model

Effectiveness may be limited by pressure drop and cost. Few suppliers

0%

0%

0%

0.025%

Tilt angle down to 10% or maximum heat rate.

Bypass valve or pump speed control over full range

Control valve over full range

Control valve or pump speed control over full range.

2.

SOURCE : ASHRAE 1996 SYSTEMS & EQUIPMENT HANDBOOK CHAPTER 42.

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APPENDIX-2

SELECTION OF HEAT RECOVERY WHEEL 1.0

Selection of Heat Recovery Wheel is normally done using manufacturer’s data. However a typical calculation procedure for HVAC system is given below for guidance.

STEP-1 Determine the Design Conditions (Indoor as well as outdoor temperature/ Humidity and Enthalpy Conditions).

STEP-2 For certain applications, as many as four different wheel sizes may be utilised. The ultimate selection is based on both performance and economic considerations. A mid point selection can be made using the following formula : Nominal wheel area (sq.m) =

Supply or Exhaust Volume (CMH) 3600 x (wheel face velocity in m/s)

Manufacturer’s data furnishes the wheel size which approximates the required face area. A typical (manufacturer’s) wheel face area Vs diameter (size) of wheel is given in APPENDIX-3. A nominal face velocity of 3 to 3.5 m/s. can be assumed to arrive at the wheel area.

STEP-3 Based on the wheel face velocity and the air flow requirements, the thermal effectiveness and pressure drop can be determined, using manufacturer’s data. A typical wheel selection chart is furnished in TABLE-2. In the absence of such data, assume a thermal effectiveness of approx. 80% (for total heat recovery) and a pressure drop of 10 to 18 mm (0.4 to 0.7 inches) of water column (as given in APPENDIX-1).

STEP-4 Using the following equations, supply air and exhaust air conditions can be found out. Effectiveness = W s (X1 – X2 )/Wmin (X1 – X3 ) Effectiveness = We (X4 – X3 )/Wmin (X1 – X3 ) where Ws = Supply air volume ISSUE R0 FORM NO. 120 R1

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APPENDIX-2 (CONTD.)

We = Return air volume/ Exhaust air volume

♦

X1, 2, 3 & 4

=

Dry bulb temperatures, humidity ratios and enthalpies corresponding to

X1

=

Outdoor air (Entering)

X2

=

Supply air (Leaving)

X3

=

Return air (Entering)

X4

=

Exhaust air (Leaving)

For equal supply volume and exhaust volume Wx = Wmin

♦

= WExhaust

For unequal flow, the air flow to be divided as per the ratio = Wmax Wmin

STEP 5 a.

Calculate Performance by finding X1 , X2 , X3 & X4 : temperatures, humidity ratios and enthalpies at Entering & Leaving conditions.

b.

Calculate the potential reduction in plant capacity (Tons of Refrigeration) using, Energy recovered = Air flow x density of air x (Enthalpy in – Enthalpy Out)

2.0

EXAMPLE PROBLEM Building

:

TCS Sholinganallur – SSDC complex

Location

:

Chennai

Fresh Air quantity

:

15300 CMH (9000 CFM)

To find

:

Reduction in Refrigeration chiller capacity using HRW. ISSUE R0 FORM NO. 120 R1

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APPENDIX-2 (CONTD.) DB ° C (°° F)

WB ° C (°° F)

Absolute Humidity (kg/kg)/ (gr/lb)

Enthalpy kJ/kg (Btu/lb)

Summer

39.4 (103)

27.8(82)

0.019(133)

89 (45.9)

Monsoon

28.3(83)

26.7(80)

0.022(153)

84 (43.7)

Winter

18.3(65)

13.9(57)

0.0018(56)

39 (24.5)

24(75)

17.5(63.5) (55% RH)

0.01(71.4)

52 (29.1)

Sl No

1.

Outdoor conditions

2.

1.

Conditions

Inside Conditions

Heat Recovery Wheel Nominal Wheel area (sq.m) = 15300 CMH = 1.2 sq.m (13 sq.ft) 3600 x 3.5 where face air velocity across the wheel = 3.5 m/s (700 FPM)

2.

From Manufacturer’s Performance data (TABLE-2)/APPENDIX-2. For a face velocity of 3.5 m/s (700 FPM). Model TE3-13 will handle 15580 CMH (9170 CFM) and the size of wheel = 1930 mm x 1930 mm (76” x 76”) Actual wheel area = 1.2 sq.m (13.1 sq.ft)

3.

Effectiveness for the model selected = 78.5% This can also be found out from manufacturer’s charts/software

4.

Performance Ws

=

Wmin =

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 15300 Using Equation (3) given in clause 5.0 X 2 = 39.4 +  (24 − 39.4)  0.785  15300  = 39.4 – 12.22 = 27.18 Say 27.2°C (81°F) i.e Fresh air (leaving HRW) DBT = 27.2°C (81°F) Similarly X2 (W)

X2 (H)

=

0.019 – 0.785 (0.019 – 0.01)

=

0.019 - 0.0071 = 0.0119 Say 0.012 kg/kg of dry air (84 gr/lb)

=

89 - 0.785 (89-52)

=

59.955 say 60 kJ/kg (32.7 Btu/ lb)

Similarly for exhaust air conditions, using equation (4) as given in clause 5.0.  15300 X4 = 24 − 0.785 (24 − 39.4 ) = 36.1°C = 97°F  15300  i.e. Exhaust air (leaving) DBT = TDB = 36.1 °C X4 (W)

= =

0.010 + 0.785 (0.019 - 0.010) 0.0171 kg/kg of dry air (119.8 gr/lb)

X4 (H)

=

52 + 0.785 (89-52) = 81.045 kJ/kg

Similar calculations Sh. 19 of 28. 5.

can be made for Monsoon and Winter conditions - See

Energy Recovered considering supply air entering and leaving conditions. Summer

= 4.5 x 9000 x (45.9 – 32.7) = 15300 x 1.2 x (89-60) = 147.9 kW = 42 TR

This is also equal to : 15300 x 1.2(81.1 - 52) = 1484.4 kW or 42.2 TR., considering exhaust air entering and leaving conditions.

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APPENDIX-2 (CONTD.)

SITE DATA ABSOLUTE HUMIDITY

ENTHALPY

kg/kg OF DRY AIR (GR/LB)

kJ/kg (BTU//LB)

DBT ºC (°° F)

WBT ºC (°° F)

SUMMER

39.4(103)

27.8(82)

0.019 (133)

89.0(45.9)

MONSOON

28.3(83)

26.7(80)

0.022 (153)

84.0(43.7)

WINTER

18.3(65)

13.9(57)

0.008(56)

39.0(24.5)

24(75)

17.5°C(63.5°) 55 % RH

0.01(71.4)

52.0(29.1)

OUTSIDE COND.

INSIDE COND. FRESH AIR QTY

15300 CMH

(9000 CFM)

DESIGN PARAMETERS (From TABLE-2) DESIGN EFFECTIVENESS

:

78.5%

FACE VELOCITY

:

3.5 m/sec (700 FPM)

WHEEL DIA SELECTED

:

1930 mm (76")

PR. DROP SUPPLY

:

17 mm wc (0.67”)

PR. DROP EXHAUST

:

17 mm wc (0.67”)

WHEEL MODEL SELECTED

:

HRW TE3-13 OF ECO-FRESH (ARCTIC INDIA)

NOTE : DBT WBT

: :

DRY BULB TEMPERATURE (ºC) WET BULB TEMPERATURE (ºC) ISSUE R0 FORM NO. 120 R1

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APPENDIX-2 (CONTD.) HEAT RECOVERY WHEEL MODEL NO.

HRW –1800

SUPPLY AIR ENTERING AIR QTY : 15300 CMH + (Purge air) [(9000 CFM + Purge Air)] DBT WBT ABS. HUMIDITY °C (°° F) ° C (°° F) kg/kg (gr/lb)

ENTHALPY kJ/kg

(Btu/1b)

SUMMER

39.4

(103)

27.8 (82)

0.019

(133.0)

89.0

(45.9)

MONSOON

28.3

(83)

26.7 (80)

0.022

(153.0)

84.0

(43.7)

WINTER

18.3

(65)

13.9 (57)

0.008

(56.0)

39.0

(24.5)

SUPPLY AIR LEAVING AIR QTY : 15300 CMH (9000 CFM) °C

DBT (°° F)

WBT ° C (°° F)

ABS. HUMIDITY

ENTHALPY

kg/kg

(gr/lb)

kJ/kg

(Btu/1b)

SUMMER

27.2

(81)

20.5

0.012

(84.6)

60.0

(32.7)

MONSOON

25

(77)

20.0

0.0127

(88.9)

59.0

(32.3)

WINTER

22.8

(73)

17.2

0.0097

(68.1)

47.5

(28.1)

EXHAUST AIR LEAVING AIR QTY : 15300 CMH + Purge Air [(9000 CFM + (Purge Air))] °C

DBT (°° F)

WBT ° C (°° F)

ABS. HUMIDITY kg/kg

(gr/lb)

ENTHALPY kJ/kg

(Btu/1b)

SUMMER

36.1

(97)

26

0.017

(1119.8)

81.1

(42.3)

MONSOON

26.7

(80)

24.8

0.019

(135.5)

76.0

(40.6)

WINTER

19.4

(67)

14.8

0.0085

(59.3)

43.5

(25.5)

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APPENDIX-2 (CONTD.) EXHAUST AIR ENTERING AIR QTY : 15300 CMH (9000 CFM) DBT WBT °C (°° F) ° C (°° F) SUMMER

23.9

(75)

MONSOON

23.9

(75)

WINTER

23.9

(75)

18°C (64.4) 18°C (64.4) 18°C (64.4)

RH%

ABS. HUMIDITY

ENTHALPY

kg/kg

(gr/lb)

kJ/ kg

(Btu/lb)

55

0.010

(71.4)

52

(29.1)

55

0.010

(71.4)

52

(29.1)

55

0.010

(71.4)

52

(29.1)

= = =

42 TR 36.2 TR 43.4 kW

ENERGY RECOVERED SUMMER MONSOON WINTER

: : :

147.9 kW 127.4 kW -

or or -

NOTE 1 : The calculations shown above are typical and values approximate for the purposes of illustration. Similar procedure could be used for preliminary estimation. Actual reduction in Refrigeration Capacity (or Heating Capacity) or energy recovered shall be obtained by Equipment Vendor. NOTE 2 : Purge air quantities will be determined by manufacturer. NOTE 3 : Calculation of Enthalpy Enthalpy values could be obtained from a Psychrometric chart. However Enthalpy Values can be calculated. ASHRAE Fundamentals Volume (S.I. Units) 1997 gives formulae for calculation of Enthalpy for a given Dry Bulb Temperature (ºC) and Humidity Ratio (Kg/Kg of dry air) (Refer Table 1 in Chapter 6 of ASHRAE FUNDAMENTALS-1997) IN S.I. UNITS Moist air enthalpy = t + w (2501 + 1.805t) kJ/kg of dry air Where t = dry bulb temperature (saturated air) W = Humidity Ratio.

--

(1)

 Pw  W can be calculated using W = 0.622   P − Pw  Where Pw = partial pressure of water vapour at tºC (kPa) and P = Atmospheric pressure (kPa).

ISSUE R0 FORM NO. 120 R1

TCE CONSULTING ENGINEERS LIMITED DESIGN GUIDE FOR HEAT RECOVERY WHEEL

TCE.M6-ME811-314

SECTION: TABLE-2

SHEET 22 OF

26

APPENDIX-2 (CONTD.) TABLE-1 TYPICAL MANUFACTURER’S PERFORMANCE DATA HEAT RECOVERY WHEEL (HRW) Velocity

Effic.

Press Drop

FPM M/S

%

IN MM

MODEL NUMBERS MODEL/

TE303

TE305

TE309

TE313

TE318

TE324

TE328

TE335

TE343

TE346

TE3-56

TE3-70

2.8

5.3

8.6

13.1

18.1

23.8

28.3

35.2

42.9

46.4

56

70.4

840 1430

1590 2700

2580 4380

3030 6680

5430 9230

7140 12130

8490 14430

10560 17940

12870 21870

13920 23650

16800 28540

21120 35880

300 1.53

88.0

0.29 7.4

AREA (SQ.FT) SCFM CMH

400 2.03

86.0

0.37 9.4

SCFM CMH

1120 1900

2120 3600

3440 5850

5240 8900

7240 12300

9520 16170

11320 19230

14080 23920

17160 29160

18560 31530

22400 38060

28160 47840

500 2.54

82.5

0.45 11.4

SCFM CMH

1400 2380

2650 4500

4300 7310

6550 11130

9050 15380

11900 20220

14150 24040

17600 29900

21450 36440

23200 39420

28000 47570

35200 59810

600 3.05

80.5

0.56 14.2

SCFM CMH

1680 2850

3180 5400

5160 8770

7860 13350

10860 18450

14280 24260

16980 28850

21120 35880

25740 43730

27840 47300

33600 57090

42240 71770

700 3.56

78.5

0.67 17.0

SCFM CMH

1960 3330

3710 6300

6020 10230

9170 15580

12670 21530

16660 33660

19810 33660

24640 41860

30030 51020

32480 55180

39200 66600

49280 83730

800 4.06

77.0

0.79 20.1

SCFM CMH

2240 3810

4240 7200

6880 11690

10480 17810

14480 24600

19040 32350

22640 38470

28160 47840

34320 58310

37120 63070

44800 76120

56320 85690

900 4.57

76.0

0.94 23.9

SCFM CMH

2520 4280

4770 8100

7740 13150

11790 20030

16290 27680

21420 36390

25470 43270

31680 53820

38610 65600

41760 70950

50400 85630

63360 107650

1000 5.08

74.5

1.05 26.7

SCFM CMH

2800 4760

5300 9010

8600 14610

13100 22260

18100 30750

23800 40440

28300 48080

35200 59800

42900 72890

46400 78830

56000 95140

70400 119610

1100 5.59

73.5

1.18 30.0

SCFM CMH

3080 5230

5830 9910

9460 16070

14410 24480

19910 33830

26180 44480

31130 52890

38720 65790

47190 80180

51040 86720

61600 104660

77440 131570

DIMENSIONS AND WEIGHTS Height and Width ‘A’

In mm

40 1016

52 1321

64 1626

76 1930

88 2235

100 2540

112 2845

124 3150

136 3454

141 3581

154 3912

172 4369

Depth in Flow Direction “B”

In mm

19 483

19 483

19 483

19 533

21 533

21 533

21 533

23 584

23 584

23 584

23 584

23 584

Net Weight

lb Kg

380 173

520 236

700 318

1070 486

1300 590

1640 745

2640 1199

3000 1362

3390 1540

3570 1621

4030 1830

4680 2125

SOURCE : ECO-FRESH HEAT WHEEL BY ARCTIC INDIA (FOR REFERENCE ONLY)

ISSUE R0 FORM NO. 120 R1