Calculations for Merchant Ship HVAC Design (2015).pdf

Technical and Research Program The Society of Naval Architects and Marine Engineers 99 Canal Center Plaza, Ste 310, Alexandria, VA 22314 www.sname.org

T&R Bulletin

4-16

Recommended Practices for Ship Heating, Ventilation & Air Conditioning Design Calculations

(2015)

SNAME Technical and Research Bulletin 4-16 (2015 Interim)

Recommended Practices for Merchant Ship HVAC Design Calculations ii iii

TECHNICAL AND RESEARCH BULLETIN 4-16 (2015)

RECOMMENDED PRACTICES FOR SHIP HEATING,

VENTILATION & AIR CONDITIONING DESIGN

CALCULATIONS

Prepared by

Raja Awwad

Senior Principal Mechanical Engineer

HVAC Section Head

Gibbs & Cox, Inc.

Kevin Prince

Assistant Vice President, Engineering

Gibbs & Cox, Inc.

And

Nicholas DeLuca Mark Gates Michelle Gahagan

Marine Engineer Marine Engineer Marine Engineer

Gibbs & Cox, Inc. Gibbs & Cox, Inc. Gibbs & Cox, Inc.

In conjunction with the

SHIPS MACHINERY COMMITTEE

December 2015

Published by

The Society of Naval Architects and Marine Engineers 99 Canal Center Plaza, Alexandria, Virginia 22314

Copyright 2015 by the Society of Naval Architects and Marine Engineers with rights reserved.

iv

This Bulletin was prepared under direction from the Ships Machinery Committee for

THE SOCIETY OF NAVAL ARCHITECTS AND MARINE ENGINEERS

TECHNICAL AND RESEARCH PROGRAM

Mr. Frederick (“Rick) H. Ashcroft, Chairman

Reviewed and Approved by:

SHIPS MACHINERY COMMITTEE

Mr. Richard Delpizzo, Committee Chair

Mr. Robert J. Bazzini Mr. Richard D. Hepburn

Mr. Robert S. Behr Mr. Bahadir Inozu

Mr. John W. Boylston Mr. Charles A. Narwicz

Mr. William G. Bullock Mr. Mark F. Nittel

Mr. Hannon Marshal Burford Mr. Michael G. Parsons

Mr. Allen Chin Mr. Kevin Prince

Mr. Joseph H. Comer Mr. Michael J. Roa

Mr. W. Mark Cummings Mr. David R. Rodger

Mr. John J. Dumbleton Mr. Alan L. Rowen

Mr. Jose Femenia Mr. Peter George Schaedel

Mr. Earl W. Fenstermacher Mr. John Thomas Schroppe

Mr. Robert M. Freeman Mr. William J. Sembler

Mr. Rob Geis Mr. Tony Teo

Mr. Joseph D. Hamilton Mr. Richard P. Thorsen

Mr. Richard W. Harkins Mr. Ivan Zgaljic

Mr. John F. Hennings v

It is understood and agreed that nothing expressed herein is intended or shall be construed to give any person, firm, or corporation any right, remedy, or claim against SNAME or any of its officers or members.

vi

Preface

The revised Technical and Research Bulletin No. 4-16 has been prepared to standardize heating, ventilation, and air conditioning calculations for merchant and naval ship designs. The revised Bulletin introduced calculations unique to naval ships but can be applied to merchant ships. There is a new section dedicated to the HVAC conversion factor and a section on humidity control. The Bulletin has a new section on ship’s air balance, demonstrated with examples in the cooling and heating seasons that air balance goes beyond supplying and exhausting the same volume of air. All sample forms were revised and included in the last section. Two examples were included to demonstrate the difference between a “push through” and a “pull through” air conditioning systems with demonstrative psychrometric charts. There is no implication of warranty by The Society of Naval Architects and

Marine Engineers that use of this guideline will ensure successful performance of vessels and/or machinery including compliance with contract specifications, Regulatory bodies or classification societies.

Acknowledgements

Many thanks to Mr. Richard Delpizzo of American Bureau of Shipping for presenting the authors with the opportunity to revise T&R Bulletin 4-16.

The principal author, Raja Awwad, was fortunate to have a group of talented engineers at Gibbs

& Cox, Inc. to assist in this endeavor: Nicholas Deluca, Mark Gates and Michelle Gahagan.

I am equally grateful for the SNAME T&R Steering Committee Chair, Mr. Rick Ashcroft, SNAME T&R Coordinator, Mr. Alex Landsburg and Mr. Anthony Maples Owner and Founder of Maples Environment Services, LLC for their review and insightful comments.

Thanks to Mr. Kevin Prince for his support and valuable contribution.

Last but not least, I extend my appreciation to: Jessica Baker, Senior Mechanical Engineer, Preconstruction Manager, Southland Industries and Patricia McGinn, Engineering Group

Assistant Manager Machinery Department, Gibbs & Cox, Inc. for their review and comments.

Table of Contents

PREFACE .................................................................................................................................... VI

ACKNOWLEDGEMENTS ....................................................................................................... VI

DEFINITIONS

SYMBOLS

1.0 INTRODUCTION

1.1 GENERAL

1.2 PECULIARITIES OF MARINE HEATING, VENTILATING, AND AIR CONDITIONING DESIGN

1.3 SCOPE

2.0 DESIGN CONDITIONS

2.1 AIR CONDITIONED SPACES

2.1.1 Air Conditioning Design Temperatures

2.1.2 Heating Design Temperatures

2.1.3 Minimum Outdoor Air (Replenishment) Requirements

2.1.3.1 Occupancy

2.1.3.2 Infiltration

2.2 NON-AIR CONDITIONED SPACES

2.3 AIR CONDITIONING SYSTEM DESIGN PRACTICE

2.3.1 Supply Air

2.3.2 Relative Humidity

2.4 AIR AND WATER VAPOR

3.0 ROOM LOAD COMPONENTS AND CALCULATIONS

3.1 TRANSMISSION LOAD

3.1.1 Transmission Heat Gain (Loss)

3.1.2 Sample Calculations – Transmission Only

3.2 SOLAR + TRANSMISSION LOAD

3.2.1 Solar + Transmission Heat Gain

3.2.2 Sample Calculations – Transmission Only

3.3 LIGHTING LOAD

3.3.1 Lighting Heat Gain

3.3.2 Sample Calculations

3.4 EQUIPMENT LOAD

3.4.1 Equipment Heat Gain

3.4.2 Direct Return from Electronics / Electrical Equipment

3.4.3 Sample Calculations

3.5 PERSONNEL LOAD

3.5.1 Personnel Heat Gain

3.5.2 Sample Calculations

3.6 HVAC CONVERSION FACTOR

3.7 INFILTRATION LOAD

3.7.1 Infiltration Heat Gain (Loss)

3.7.2 Sample Calculations

3.8 HUMIDITY CONTROL

3.8.1 Constant Summer Reheat Applications

3.8.2 Sample Calculations

3.9 VENTILATION LOAD

3.9.1 Allowable Temperature Rise Calculations

3.9.1.1 Transmission Load

3.9.1.2 Solar Load

3.9.1.3 Lighting Load

3.9.1.4 Equipment Load

3.9.2 Rate of Air Change

3.9.3 Precooling

3.9.4 Sample Calculations

4.0 SYSTEM COMPONENTS AND CALCULATIONS

4.1 ROOM LOAD

4.2 FAN LOAD

4.2.1 Fan Motor Heat Gain

4.2.2 Sample Calculations

4.3 SUPPLY DUCT LOAD

4.3.1 Supply Duct Heat Gain

4.3.2 Sample Calculations

4.4 RETURN PATH LOAD

4.4.1 Return Path Heat Gain

4.4.2 Sample Calculations

4.5 OUTDOOR AIR LOAD

4.5.1 Occupancy Calculations

4.5.2 Sample Calculations

5.0 SHIP’S AIR BALANCE

6.0 HVAC CALCULATION BOOKLET

6.1 COMPARTMENT EQUIPMENT LIST

6.1.1 Sample Forms

6.2 HEATING AND COOLING LOAD CALCULATIONS

6.2.1 Sample Forms

6.3 PSYCHROMETRIC CHART

6.4 COOLING ANALYSIS

6.5 HEATING ANALYSIS

6.6 SUPPLY SYSTEM ANALYSIS

6.7 EXHAUST SYSTEM ANALYSIS

BIBLIOGRAPHY/REFERENCES

APPENDIX A HVAC CASE STUDY

DEFINITIONS

Air conditioning: The process of treating air to control temperature and relative humidity.

Cooling season: The time of the year when cooling is required.

Heating season: The time of the year when heating is required.

High occupancy spaces: Spaces having twelve (12) or more occupants.

Ventilation: The supply and exhaust of outside air by fans for the purpose of maintaining temperatures, and removing offensive odors and contaminants from spaces.

Outdoor air: Weather air supplied to the ship by ventilation fans to ventilated spaces, as replenishment to air conditioning systems, or as infiltration into the ship.

Infiltration: Uncontrolled leakage of outside air into an air conditioned space.

Direct return: Dedicated return or exhaust terminal located within six inches of the discharge ventilation opening of equipment with an internal ventilation blower.

Pull-through system: An air conditioning system where the fan is located downstream of the cooling coil.

Push-through system: An air conditioning system where the fan is located upstream of the cooling coil.

Constant summer reheat: Sensible heat added for the purpose of controlling relative humidity in spaces with high latent (moisture) loads.

Pressurized space: An air conditioned space where, by design, the amount of replenishment air exceeds the amount of exhaust.

Sensible heat factor: Is the ratio of room total sensible heat load to the room total load.

SYMBOLS

A – area of deck or room boundary [ft

] ql – latent heat [Btu/hr]

B.F. – ballast factor qs – sensible heat [Btu/hr] cp,a – specific heat of dry air [Btu/lb/°F] ql,d – equipment latent heat dissipation [Btu/hr] cp,w – specific heat of water vapor [Btu/lb/°F] qs,d – equipment sensible heat dissipation [Btu/hr]

CSR – constant summer reheat ql,p – personnel latent heat dissipation [Btu/hr/person]

F – HVAC conversion factor [Btu-lb-min/hr/ft /°F] qs,p – personnel sensible heat dissipation [Btu/hr/person]

F.W. – fluorescent bulb wattage [W] SHF – sensible heat factor

Gsf – glass solar factor [Btu/hr/ft ] ∆T – temperature difference [°F] ho – enthalpy of outdoor air [Btu/lb of dry air] ∆Te – solar effective temperature difference [°F] hr – enthalpy of room air [Btu/lb of dry air] Tf – temperature rise due to fan motor heat [°F]

H.F. – hood factor To – temperature of outdoor air [°F]

I.W. – incandescent bulb wattage [W] Tp – design temperature of precooled supply air [°F]

L.C. – load constant [Btu/hr/ft ] Tr – temperature rise [°F]

LAT – coil leaving air temperature [°F] Ts – space design air temperature [°F]

Of – minimum outdoor air rate [CFM/person] U – overall coefficient of heat transfer [Btu/hr/ft /°F]

P – number of room occupants [persons] U.F. – equipment use factor

Q – air flow quantity [CFM] V – room volume [ft

QCC – air flow through the precooling coil [CFM] v – specific volume [ft /lb] q – total heat [Btu/hr] W – humidity ratio [lb of water vapor/lb of dry air] qf – fan motor heat [Btu/hr] ηmotor – motor efficiency

1.0 INTRODUCTION

1.1 General

The primary functions of marine heating, ventilation, and air conditioning (HVAC) systems are to provide comfort and healthy conditions for the crew and passengers, to maintain satisfactory operation of equipment, prevent build up of offensive odors and dangerous gases, and prevent spoilage of supplies by controlling ambient conditions.

This is accomplished by establishing the heating, ventilating, and air conditioning requirements for each space and then preparing a set of "load" calculations and analyses to determine the sizes of HVAC equipment that meet all established requirements.

1.2 Peculiarities of Marine Heating, Ventilating, and Air

Conditioning Design Basic design considerations are similar for land-based and marine HVAC installations; however, there are a number of factors uncommon to land-based installations which must be considered when designing a marine HVAC system.

A marine HVAC installation:

• Is subjected to forces induced by pitch, roll, and slamming of the ship.

• Is designed to operate over a wide range of weather conditions and can be subjected to extreme changes in weather conditions (temperature, humidity and solar) within a short period of time.

• Is designed to function in a corrosive seawater environment.

• Is relatively compact due to space limitation on ships.

• Is more susceptible to noise problems because the enclosed areas are relatively small, close to equipment and machinery spaces, and the ship’s structure affords very little inherent sound dampening.

• Must be very rugged because ship operating schedules severely limit the time available for maintenance by specialized land-based personnel.

1.3 Scope

This bulletin contains guidance to assist in designing a marine HVAC installation for a merchant ship. Detailed room and system design criteria are provided for all spaces of a typical merchant ship, with the exception of the engine room (machinery space) and cargo spaces, which, due to the complexity of equipment, heat liberation, and commodity ventilation requirements, are not within the scope of this bulletin. The instructions include the detailed calculations required to determine both room and system loads. Typical HVAC calculations and forms are included to clarify the application of the data contained herein.

2.0 DESIGN CONDITIONS

A ship's HVAC requirements are dependent on specific data which must be gathered before the necessary calculations can be made. This data includes:

• Outside air design dry bulb (DB) and wet bulb (WB) temperatures for the cooling season.

• Outside air design dry bulb temperature for the heating season.

• Room design dry bulb temperatures in the cooling and heating seasons.

• Room design relative humidity (RH) for the cooling season.

• Design seawater temperature for both the cooling and heating seasons.

• Ship’s general arrangement drawing.

• Boundary constructions for each space.

• Lights in each space.

• Equipment in each space.

• Occupancy in each space.

• Room ventilation requirements.

• Type of cooling system.

• Type of heating system.

• Type of ventilation system.

Most of the above design data is listed in, or can be determined from, information contained in the ship's construction specifications, general arrangement plans, joiner arrangement plans, structural plans, lighting arrangement plans, and equipment lists; however, where the specifications and plans are silent on design data, recommended values in this section should be used until the data becomes available.

2.1 Air Conditioned Spaces

Normally, all staterooms, cabins, lounges, recreation rooms, mess rooms, dining rooms, offices, slop chests, radio and radar transmission rooms, electronic rooms, Gyro rooms, steering gear rooms, chart rooms, hospitals, wheelhouses and passages within A/C areas shall be cooled during the cooling season and heated during the heating season. Stairwells within the air conditioned envelope shall be indirectly conditioned by return air, wherever possible.

2.1.1 Air Conditioning Design Temperatures

Outside Air Dry Bulb 95°F

Outside Air Wet Bulb 82°F

Inside Air Dry Bulb 78°F

Inside Air Wet Bulb 66.5°F (corresponds to 55% RH)

Seawater 85°F

2.1.2 Heating Design Temperatures

Outside Air Dry Bulb 0°F

Inside Air Dry Bulb 70°F

Medical Spaces Dry Bulb 75°F

Seawater 28°F

2.1.3 Minimum Outdoor Air (Replenishment) Requirements

The minimum outdoor air to be supplied to an air conditioning system shall be determined by room occupancy.

2.1.3.1 Occupancy

Minimum outdoor air based on room occupancy shall be 12 CFM/person for high occupancy spaces (e.g., recreation rooms, dining rooms and mess rooms) and 15 CFM/person for low occupancy air conditioned spaces (e.g., staterooms and offices).

These rates shall only be increased to achieve air balance within each subdivision, but shall not exceed 50

CFM/person.

2.1.3.2 Infiltration

For spaces such as the Wheelhouse (Pilot House) with doors that open directly to the weather, minimum outdoor air rate of change shall be 60 minutes, but shall not exceed 50 CFM/person.

2.2 Non-Air Conditioned Spaces

Spaces which are not air conditioned shall be heated and/or ventilated in accordance with the requirements of Table

2-1. Non-air conditioned spaces not mentioned in the table shall be treated the same as similar spaces listed. In all cases, the allowable temperature rise must provide room ambient temperatures within the temperature ratings of electrical equipment installed in the room. The design temperature rises in Table 2-1 are based on outside air and seawater design temperatures noted in Section 2.1.1. These temperature rises shall be adjusted if different outside air and seawater temperatures are used.

Table 2-1: Ventilation & Heating Recommendations for Non-Air Conditioned Spaces

Type Space

Ventilation Heating

Temperature

Rise, Tr (°F)

Max. Air Change, R/C (minutes)

Temperature, Ts (°F)

Air Cond. Equip. and Fan Rooms 15 - 40

Baggage 15 15 60

Battery Room - 4 40

Bonded Stores 15 10 40

Bosn’s Stores - 20 -

Bow Thruster Room 15 6 40

Butcher Shop 10 4 40

Canvas Room - 6 40

Capstan Machinery Room 15 6 -

Cargo Control Room - 10 70

Carpenter Shop 10 6 60

Cleaning Gear Locker - 4 -

CO2 Room 15 6 -

Deck Lockers, Deck Stores 15 - -

Deck Toilet - 6 60

Dry Stores - 15 -

Elec. & Engrs. Workshops 10 6 60

Elev. Mach. Room 15 6 -

Emerg. Gen. Room 15 10 40

Engrs. Stores 20 30 -

F.O. Filling Station - 6 -

Galley and Galley Pantry * 10 1 50

Table 2-1 Continued

Type Space

Ventilation Heating

Temperature

Rise, Tr (°F)

Max. Air Change, R/C (minutes)

Temperature, Ts (°F)

Garbage Room 10 1 -

Hawser Room - 20 -

Lateral Thruster Comp’t 15 6 -

Laundries – Service ** 10 4 70

Linen Lockers, Daily Service - 15 -

Linen Room – Clean 10 5 60

Linen Room – Soiled 10 4 -

Machine Shop 15 6 60

Oil Skin Lockers - 4 -

Paint Rooms, Deck & Engr’s. 10 4 -

Pantries Service 10 4 -

Passageway, Outside A/C Spaces 10 6 -

Pump Room 15 3 40

Sewage Tank Space 15 6 -

Shaft Alley - 10 -

Stores – Ship’s Dry Daily Galley

& Bulk Food 10 4 50

Stores – Steward’s & Misc. 10 6 -

Stevedores Toilets - 6 60

Toilet & Showers 10 4 70

Winch M.G. 15 6 -

Windlass Machinery Room 15 6 40

* The minimum exhaust air from a Galley shall be equal to the total airflow from all hoods in the Galley.

** The minimum exhaust air from a Laundry space shall be equal to the clothes dryers’ total blower airflow.

2.3 Air Conditioning System Design Practice

The following criteria are considered to be good design practices. It is recognized that particular installations may require the use of values other than those listed below. It is recommended, however, that when other values are used, their use be considered on a case by case basis only. It is noted that example cases used throughout this bulletin are based on outside air and seawater design temperatures noted in Section 2.1.

2.3.1 Supply Air

The maximum design temperature difference between the supply air at the room terminal and the room design temperature shall be 30°F.

Room supply air shall be based on the room sensible load, supply air dry bulb, and room design dry bulb temperatures.

2.3.2 Relative Humidity

• The design off-coil relative humidity of supply air shall not exceed 95%.

• The system design relative humidity shall not exceed 55%.

• Individual rooms within a system shall not exceed 55% RH.

2.4 Air and Water Vapor

The air and water vapor for air conditioning calculations shall be considered as having the following physical properties:

Dry Air Specific Heat 0.24 Btu per lb. per o F

Dry Air Specific Volume 13.34 ft per lb.

Water Vapor Specific Heat 0.45 Btu per lb. per o F

Heat of Vaporization 1055 Btu per lb.

Physical properties of air and water vapor for ventilation, infiltration, and preheating are discussed in Section 3.6 of this bulletin.

3.0 ROOM LOAD COMPONENTS AND

This section discusses the first phase in designing an HVAC system which is to establish the cooling, heating, and ventilation load requirements for each individual room.

Table 3-1 lists room load components, the section which describes each component, and when each component is considered.

Table 3-1: Room Load Components

Component Section

Cooling

Calculations

Heating

Calculations

Ventilation

Calculations

Transmission 3.1 X X X

Solar 3.2 X - X

Lights 3.3 X X X

Equipment 3.4 X - X

Personnel 3.5 X - -

Infiltration 3.7 - X -

Ventilation 3.9 - - X

Production schedules may require that HVAC load calculations be made concurrent with or prior to the completion of design development and the selections of equipment and lighting facilities. This requires the HVAC engineer to make assumptions (e. g., boundary constructions, lighting requirements, and equipment sizes). Calculations based on such assumptions must be considered as estimates and must be revised when the actual conditions and requirements have been established to ensure the installed HVAC system is designed to meet actual conditions and requirements.

3.1 Transmission Load

Transmission load is the sensible heat flow through a boundary due to the temperature differential across the boundary surfaces.

3.1.1 Transmission Heat Gain (Loss)

Transmission heat gain or loss for a boundary is calculated using the equation:

TAUq ∆××= (3-1)

Where ∆T is the temperature difference across the boundary.

Separate calculations must be made for each boundary or portion which has a different U or ∆T value.

Where rooms have composite boundaries of steel and joiner work, the boundary areas shall be calculated using dimensions which extend to the steel. Molded (steel to steel) dimensions shall be used in Equation 3-1.

Dimensions shall be rounded off to the nearest foot.

When a room design temperature is not specified, the temperatures listed in Table 3-2 shall be used or an abbreviated heat balance calculation may be used to estimate the room temperature.

Values of U shall be taken from the current edition of Thermal Insulation Report, SNAME Technical and

Research Bulletin No. 4-7.

For cooling load calculations, the cooling effect of an adjacent space is not considered unless the lower temperature is maintained by air conditioning or refrigeration equipment. For heating load calculations, the heat gain through a boundary is deducted only if the higher temperature of the adjacent space is maintained by a heating system.

Table 3-2: Transmission Calculation Temperatures for Non-Air Conditioned Spaces

Type of Space Ts – Cooling (°F) Ts – Heating (°F)

Air Conditioning Equipment and Fan Rooms 110 40

Baggage 110 60

Battery Room 110 40

Bonded Stores 110 40

Bosn’s Stores 115 30

Bow Thruster Room 110 40

Table 3-2 Continued

Butcher Shop 105 40

Canvas Room 105 40

Capstan Machinery Room 110 30

Cargo Control Room 105 70

Cargo Oil Tanks 140 40

Cargo Spaces Above Waterline 120 20

Cargo Spaces Below Waterline 100 40

Carpenter Shop 105 60

Cleaning Gear Locker 105 60

CO2 Room 110 40

Deck Lockers, Deck Stores 110 30

Deck Toilet 110 60

Dry Stores 110 50

Elec. & Engrs. Workshops 105 60

Elevator Machinery Room 110 40

Emerg. Generator Room 110 40

Engrs. Stores 115 30

Fresh Water Tanks 120 35

F.O. Filling Station 110 30

F.O. Settling Tanks 120 70

F.O. Storage Tanks 120 40

Galley and Galley Pantry 105 50

Garbage Room 105 50

Hawser Room 115 30

Lateral Thruster Comp’t. 110 40

Laundries – Service 105 70

Linen Lockers, Daily Service 105 60

Linen Room – Clean 105 60

Linen Room – Soiled 105 40

Machine Shop 110 60

Machinery & Boiler Casings 140 70

Machinery & Boiler Spaces 130 50

M.G. & Resistor Rooms 110 30

Oil Skin Lockers 105 40

Paint Rooms, Deck & Engr.’s 105 40

Pantries Service 105 60

Stairwells Inside A/C Areas 80 65

Passageways & Stairwells Outside A/C Areas 105 60

Pump Room 110 40

Sewage Tank Space 110 40

Shaft Alley 105 40

Stores – Ship’s Dry Daily Galley & Bulk Food 105 50

Stores – Steward’s & Misc. 105 50

Stevedores Toilets 105 60

Toilets Outside A/C Spaces 105 60

Toilets & Showers, Medical Spaces 85 75

Toilets & Showers Within A/C Spaces 85 70

Voids Above Waterline 120 20

Voids Below Waterline 100 40

Winch M.G. Room 110 30

Windlass Machinery Room 110 40

3.1.2 Sample Calculations – Transmission Only

For cooling calculations, heat flow into a room is positive (load) and heat flow out of a room is negative (credit, is deducted from the room cooling load). For heating calculations, heat flow out of a room is positive (load) and heat flow into a room is negative (credit, is deducted from the room heating load).

CASE 1 Stateroom – deck over: Type 10 construction, molded dimensions 11'-6" x 12'-8", adjacent to an air conditioned cabin.

Cooling Calculation

∆T = 78 – 78 = 0°F (Sec. 2.1.1)

No Heat Flow q = 0

Heating Calculation

∆T = 70 – 70 = 0°F (Sec. 2.1.1)

CASE 2 Stateroom – deck below: Type 55 construction, molded dimensions 11'-6" x 12'-8", adjacent to engineers’ storeroom.

∆T = 115 – 78 = 37°F (Table 3-2 & Sec. 2.1.1)

A = 12' × 13' = 156 ft

(Sec. 3.1.1)

U = 0.11 Btu/hr/ft /°F (T&R 4-7, Type 55) q = 0.11 × 156 × 37 = 635 Btu/hr (Equation 3-1)

∆T = 70 – 30 = 40°F (Table 3-2 & Sec. 2.1.2)

(Sec. 3.1.1)

U = 0.09 Btu/hr/ft /°F (T&R 4-7, Type 55) q = 0.09 × 156 × 40 = 562 Btu/hr (Equation 3-1)

CASE 3 Stateroom – forward bulkhead: Type 6 construction, molded dimensions 8'-5" x 12'-8", adjacent to laundry.

∆T = 105 – 78 = 27°F (Table 3-2 & Sec. 2.1.1)

A = 8' × 13' = 104 ft

(Sec. 3.1.1)

U = 0.313 Btu/hr/ft /°F (T&R 4-7, Type 6) q = 0.313 × 104 × 27 = 879 Btu/hr (Equation 3-1)

∆T = 70 – 70 = 0°F (Table 3-2 & Sec. 2.1.2)

CASE 4 Stateroom – after bulkhead: Type 5 construction, molded dimensions 8'-5" x 12'-8", adjacent to stateroom toilet & shower (5') and hospital (7'-8").

T & S ∆T = 85 – 78 = 7°F (Table 3-2 & Sec. 2.1.1)

T & S A = 8' × 5' = 40 ft

(Sec. 3.1.1)

T & S U = 0.376 Btu/hr/ft /°F (T&R 4-7, Type 5)

T & S q = 0.376 × 40 × 7 = 105 Btu/hr (Equation 3-1)

Hospital ∆T = 78 – 78 = 0°F (Table 3-2 & Sec. 2.1.1)

Hospital No Heat Flow

Hospital q = 0

T & S ∆T = 70 – 70 = 0°F (Table 3-2 & Sec. 2.1.2)

T & S No Heat Flow

T & S q = 0

Hospital ∆T = 70 – 75 = -5°F (Sec. 2.1.2)

Hospital A = 8' × 8' = 64 ft

(Sec. 3.1.1)

Hospital U = 0.354 Btu/hr/ft /°F (T&R 4-7, Type 5)

Hospital q = 0.354 × 64 × -5 = -113 Btu/hr (Equation 3-1)

CASE 5 Stateroom – inboard bulkhead: Type 5 construction, molded dimensions 8'-5" x 11'-6", adjacent to passageway.

∆T = 80 – 78 = 2°F (Table 3-2 & Sec. 2.1.1)

A = 8' × 12' = 96 ft

(Sec. 3.1.1)

U = 0.376 Btu/hr/ft q = 0.376 × 96 × 2 = 72 Btu/hr (Equation 3-1)

∆T = 70 – 65 = 5°F (Table 3-2 & Sec. 2.1.2)

(Sec. 3.1.1)

U = 0.354 Btu/hr/ft q = 0.354 × 96 × 5 = 170 Btu/hr (Equation 3-1)

CASE 6 Stateroom – outboard bulkhead: Type 63 construction, molded dimensions 8'-5" x 11'-6", glass area 6 ft

, fully shaded by structural overhang.

Type 63 ∆T = 95 – 78 = 17°F (Sec. 2.1.1)

Type 63 A = (8' × 12') – 6 ft = 90 ft

(Sec. 3.1.1)

Type 63 U = 0.10 Btu/hr/ft /°F (T&R 4-7, Type 63)

Type 63 q = 0.10 × 90 × 17 = 153 Btu/hr (Equation 3-1)

Glass ∆T = 95 – 78 = 17°F (Sec. 2.1.1)

Glass A = 6 ft

(Sec. 3.1.1)

Glass U = 1.13 Btu/hr/ft /°F (T&R 4-7, Window Glass, Single Pane)

Glass q = 1.13 × 6 × 17 = 115 Btu/hr (Equation 3-1)

Type 63 ∆T = 70 – 0 = 70°F (Sec. 2.1.2)

= 90 ft

(Sec. 3.1.1)

Type 63 U = 0.09 Btu/hr/ft /°F (T&R 4-7, Type 63)

Type 63 q = 0.09 × 90 × 70 = 567 Btu/hr (Equation 3-1)

Glass ∆T = 70 – 0 = 70°F (Sec. 2.1.2)

Glass A = 6 ft

(Sec. 3.1.1)

Glass U = 1.13 Btu/hr/ft

Glass q = 1.13 × 6 × 70 = 475 Btu/hr (Equation 3-1)

CASE 7 Laundry Room – deck over: Type 57 construction, molded dimensions 11'-6" x 12'-8", adjacent to air conditioned cabin.

∆T = 78 – 105 = -27°F (Table 3-2 & Sec. 2.1.1)

(Sec. 3.1.1)

U = 0.23 Btu/hr/ft /°F (T&R 4-7, Type 57) q = 0.23 × 156 × -27 = -969 Btu/hr (Equation 3-1)

CASE 8 Laundry Room – deck below: Type 53 construction, molded dimensions 11'-6" x 12'-8", adjacent to engineers’ storeroom.

∆T = 115 – 105 = 10°F (Table 3-2)

(Sec. 3.1.1)

U = 0.28 Btu/hr/ft

/°F (T&R 4-7, Type 53) q = 0.28 × 156 × 10 = 437 Btu/hr (Equation 3-1)

∆T = 70 – 30 = 40°F (Table 3-2 & Sec. 2.1.2)

(Sec. 3.1.1)

U = 0.187 Btu/hr/ft /°F (T&R 4-7, Type 53) q = 0.187 × 156 × 40 = 1167 Btu/hr (Equation 3-1)

CASE 9 Laundry Room – forward bulkhead: Type 5 construction, molded dimensions 8'-5" x 12'-8", adjacent to clean linen room.

∆T = 105 – 105 = 0°F (Table 3-2)

∆T = 70 – 60 = 10°F (Table 3-2)

(Sec. 3.1.1)

U = 0.354 Btu/hr/ft q = 0.354 × 104 × 10 = 368 Btu/hr (Equation 3-1)

CASE 10 Laundry Room – after bulkhead: Type 6 construction, molded dimensions 8'-5" x 12'-8", adjacent to stateroom.

∆T = 78 – 105 = -27°F (Table 3-2 & Sec. 2.1.1)

(Sec. 3.1.1)

U = 0.313 Btu/hr/ft /°F (T&R 4-7, Type 6) q = 0.313 × 104 × -27 = -879 Btu/hr (Equation 3-1)

CASE 11 Laundry Room – inboard bulkhead: Type 5 construction, molded dimensions 8'-5" x 11'-6", adjacent to passageway inside A/C space.

∆T = 78 – 105 = -27°F (Table 3-2)

(Sec. 3.1.1)

U = 0.376 Btu/hr/ft q = 0.376 × 96 × -27 = -975 Btu/hr (Equation 3-1)

∆T = 70 – 70 = 0°F (Table 3-2)

3.2 Solar + Transmission Load

Solar + transmission load is the sensible heat flow through the room's weather boundaries which are exposed to the sun. This load is in addition to the transmission load for the affected boundary or boundaries, but the total load

(solar + transmission) is calculated in one step.

3.2.1 Solar + Transmission Heat Gain

Solar + transmission heat gain is calculated using a variation of Equation 3-1:

eTAUq ∆××= (3-2)

Where ∆Te is the difference between the effective temperature from Table 3-3 and room design temperature.

The values of U shall be those for “Solar Radiation” in the current edition of Thermal Insulation Report, Technical and Research Bulletin No. 4-7.

The solar + transmission heat gain for glass shall be calculated using the equation:

AGq sf ×= (3-3)

Where more than one boundary of a space is exposed to the sun, a separate heat gain calculation shall be performed for each boundary or combination of boundaries and the greatest simultaneous gain shall be used for determining the solar and transmission heat load.

The dimensions to be used when calculating the boundary area are delineated in Section 3.1.1, except that vertical shaded areas shall not be included. Shaded areas shall be calculated with the sun at a 45 o angle from the horizon.

Glass solar factors and effective temperatures for single and multi-boundary calculations are listed in Table 3-3.

If a different outside design temperature is specified, the temperatures in Table 3-3 must be adjusted accordingly, upward or downward.

Table 3-3: Glass Solar Factors & Effective Temperatures

Glass Solar Factor

Single Boundary

Calculations

Multi-Boundary

Calculations

160 Btu/hr/ft

120 Btu/hr/ft

Effective

Temperatures

Outdoor Air

Dry Bulb

Temperature

Single Boundary

Calculations

Multi-Boundary

Calculations

Horizontal

Surface

Vertical

Surface

Horizontal

Surface

Vertical

Surface

90°F 140 120 125 110

95°F 145 125 130 115

100°F 150 130 135 120

105°F 155 135 140 125

110°F 160 140 145 130

3.2.2 Sample Calculations – Transmission Only

CASE 12 Stateroom – outboard bulkhead: Type 63 construction, molded dimensions 8'-5" x 11'-6", glass area 6 ft

, completely exposed to the sun. No other boundaries exposed to the sun.

Type 63 ∆Te = 125 – 78 = 47°F (Table 3-3 & Sec. 2.1.1)

= 90 ft

(Sec. 3.2.1)

Type 63 U = 0.11 Btu/hr/ft /°F (T&R 4-7, Type 63)

Type 63 q = 0.11 × 90 × 47 = 465 Btu/hr (Equation 3-2)

Glass Gsf = 160 Btu/hr/ft (Table 3-3)

Glass A = 6 ft

(Sec. 3.2.1)

Glass q = 160 × 6 = 960 Btu/hr (Equation 3-3)

Same as Section 3.1.2 – CASE 6.

Type 63 q = 567 Btu/hr

AND

Glass q = 475 Btu/hr

CASE 13 Same as CASE 12 except the bulkhead is partially shaded by 5' structural overhang.

Cooling Calculation – Shaded Portion

Type 63 A = (5' × 12') – 6 ft = 54 ft

(Sec. 3.2.1)

Type 63 U = 0.10 Btu/hr/ft /°F (T&R 4-7, Type 63)

Type 63 q = 0.10 × 54 × 17 = 92 Btu/hr (Equation 3-1)

Glass ∆T = 95 – 78 = 17°F (Sec. 2.1.1)

Glass A = 6 ft

(Sec. 3.2.1)

Glass U = 1.13 Btu/hr/ft

Cooling Calculation – Unshaded Portion

Type 63 ∆Te = 125 – 78 = 47°F (Table 3-3 & Sec. 2.1.1)

Type 63 A = 3'5" × 12' = 3' × 12' = 36 ft

(Sec. 3.2.1)

Type 63 U = 0.11 Btu/hr/ft /°F (T&R 4-7, Type 63)

Type 63 q = 0.11 × 36 × 47 = 186 Btu/hr (Equation 3-2)

Same as Section 3.1.2 – CASE 6.

CASE 14 Stateroom – outboard bulkhead and deck above both exposed to the sun. Outboard bulkhead (same as CASE 12): Type 63 construction, molded dimensions 8'-5" x 11'-6", glass area 6 ft . Deck above: Type 62 construction, molded dimensions 11'-6" x 12'-8".

When more than one boundary may be exposed to the sun, several calculations are required to determine the maximum room load. In this case the combinations which must be considered are:

(a) Deck above solar and outboard bulkhead weather.

(b) Deck above weather and outboard bulkhead solar.

(c) Deck above solar and outboard bulkhead solar.

The combination that yields the largest load shall be used.

14(a) Cooling Calculation – deck above solar, outboard bulkhead weather.

DECK

Type 62 ∆Te = 145 – 78 = 67°F (Table 3-3 & Sec. 2.1.1)

Type 62 A = 12' × 13' = 156 ft

(Sec. 3.2.1)

Type 62 U = 0.11 Btu/hr/ft /°F (T&R 4-7, Type 62)

Type 62 q = 0.11 × 156 × 67 = 1150 Btu/hr (Equation 3-2)

BULKHEAD

= 90 ft

(Sec. 3.1.1)

Type 63 U = 0.10 Btu/hr/ft /°F (T&R 4-7, Type 63)

Type 63 q = 0.10 × 90 × 17 = 153 Btu/hr (Equation 3-1)

Glass ∆T = 95 – 78 = 17°F (Sec. 2.1.1)

Glass A = 6 ft

(Sec. 3.1.1)

Glass U = 1.13 Btu/hr/ft

TOTAL q = 1150 + 153 + 115 = 1418 Btu/hr

14(b) Cooling Calculation – deck above weather, outboard bulkhead solar.

Type 62 ∆T = 95 – 78 = 17°F (Sec. 2.1.1)

(Sec. 3.1.1)

Type 62 U = 0.09 Btu/hr/ft

Type 62 q = 0.09 × 156 × 17 = 239 Btu/hr (Equation 3-1)

BULKHEAD – Same as CASE 12.

Type 63 q = 465 Btu/hr

Glass q = 960 Btu/hr

TOTAL q = 239 + 465 + 960 = 1664 Btu/hr

14(c) Cooling Calculation – deck above and outboard bulkhead solar (use effective temperatures for multi-boundary calculations from Table 3-3).

Type 62 ∆Te = 130 – 78 = 52°F (Table 3-3 & Sec. 2.1.1)

(Sec. 3.2.1)

Type 62 U = 0.11 Btu/hr/ft

Type 62 q = 0.11 × 156 × 52 = 892 Btu/hr (Equation 3-2)

BULKHEAD

Type 63 ∆Te = 115 – 78 = 37°F (Table 3-3 & Sec. 2.1.1)

= 90 ft

(Sec. 3.2.1)

Type 63 U = 0.11 Btu/hr/ft /°F (T&R 4-7, Type 63)

Type 63 q = 0.11 × 90 × 37 = 366 Btu/hr (Equation 3-2)

Glass Gsf = 120 Btu/hr/ft (Table 3-3)

Glass A = 6 ft

(Sec. 3.2.1)

Glass q = 120 × 6 = 720 Btu/hr (Equation 3-3)

TOTAL q = 892 + 366 + 720 = 1978 Btu/hr

MAXIMUM: q = CASE 14(c) = 1978 Btu/hr

Type 62 ∆T = 70 – 0 = 70°F (Sec. 2.1.2)

(Sec. 3.1.1)

Type 62 U = 0.10 Btu/hr/ft

Type 62 q = 0.10 × 156 × 70 = 1092 Btu/hr (Equation 3-2)

BULKHEAD – Same as Section 3.1.2 – CASE 6.

CASE 15 Laundry Room – outboard bulkhead: Type 63 construction, molded dimensions 8'-5" x 11'-6", glass area 6 ft

, completely exposed to the sun.

∆Te = 125 – 105 = 20°F (Tables 3-2 & 3-3)

(Sec. 3.2.1)

U = 0.11 Btu/hr/ft q = 0.11 × 96 × 20 = 211 Btu/hr (Equation 3-2)

∆T = 70 – 0 = 70°F (Table 3-2 & Sec. 2.1.2)

(Sec. 3.1.1)

U = 0.09 Btu/hr/ft q = 0.09 × 96 × 70 = 605 Btu/hr (Equation 3-1)

3.3 Lighting Load

Lighting load is the sensible heat generated by the lights in a room. Lighting load shall be accounted for in all spaces in the cooling season, and shall be deducted from the heating load in all spaces in the heating season except berthing spaces, cabins, and staterooms.

3.3.1 Lighting Heat Gain

When the installed lighting is known, the heat gain shall be calculated for the normal room lighting mode which, for these calculations, shall be considered to be all overhead and cornice lights controlled by wall switches and work station lights; i.e., desk lights in office. The lighting load shall be calculated using the equation:

412.3...)...( ××+= FBWFWIq (3-4)

Where 3.412 is a conversion factor from Watts to Btu/hr. The ballast factor, B.F., shall be taken as 1.25.

When the installed lighting is not known, the lighting load shall be estimated using the equation:

..CLAq ×= (3-5)

The deck area, A, used when calculating estimated lighting loads shall be based on the finished dimensions, if known. Otherwise, use molded dimensions to calculate the deck area. See Table 3-4 for the appropriate value of load constant, L.C.

All estimated lighting loads must be revised when the installed lighting load is known. This is particularly important for mess rooms and other public spaces with large lighting loads.

Table 3-4: Assumed Lighting Loads (Btu/hr/ft

Deck Area)

Space Load Constant, L.C.

Passenger Staterooms 7

Captain & Chief Engineer Staterooms 7

Officer Staterooms 4

Crew Staterooms 4

Mess Rooms, Lounges & Public Spaces 9

Offices 7

Other Spaces 7

3.3.2 Sample Calculations

CASE 16 Officer Stateroom: Installed lights controlled by wall switch, three ceiling fixtures each with two

20W fluorescent bulbs. Finished dimensions, deck above: 11'-6" x 12'-2".

I.W. = 0 (Sec. 3.3.1)

F.W. = 3 × 2 × 20 = 120 Watts (Sec. 3.3.1)

B.F. = 1.25 (Sec. 3.3.1) q = 120 × 1.25 × 3.412 = 512 Btu/hr (Equation 3-4)

CASE 17 Same as CASE 16 except installed lights are not known.

Ad = 12' × 12' = 144 ft

(Sec. 3.3.1)

L.C. = 4 Btu/hr/ft (Table 3-4) q = 144 × 4 = 576 Btu/hr (Equation 3-5)

CASE 18 Officer Laundry: Installed lights controlled by wall switch, four ceiling fixtures each with two

20W fluorescent bulbs. Finished dimensions, deck above: 11'-6" x 12'-2".

I.W. = 0 (Sec. 3.3.1)

F.W. = 4 × 2 × 20 = 160 Watts (Sec. 3.3.1)

B.F. = 1.25 (Sec. 3.3.1) q = 160 × 1.25 × 3.412 = 682 Btu/hr (Equation 3-4)

3.4 Equipment Load

Equipment load is the sensible and latent heat generated by equipment operating within a space.

For air conditioned spaces, both the sensible and latent components of the equipment load are included in the cooling load calculations; however, for ventilated spaces, only the sensible component is included in the cooling load calculations.

Spaces that typically have equipment load are galleys, pantries, laundries, radio and communication rooms, wheelhouses, resistor houses, deck machinery compartments, and specialized spaces such as computer rooms or engine control rooms. Equipment load will normally be included for lounges, mess rooms, offices, and officer staterooms.

3.4.1 Equipment Heat Gain

Equipment heat gain must be based on heat dissipation data for actual equipment installed. The best source of heat dissipation data is the equipment manufacturer; however, quite often the equipment has not been selected when the HVAC calculations are performed, in which case other sources of heat dissipation data must be used, such as manufacturers' manuals for similar pieces of equipment or the ASHRAE Handbook of Fundamentals.

Table 3-5 herein provides heat dissipation data for typical marine equipment and may be used when more exact information is not available for a specific design

Table 3-5: Recommended Heat Dissipation for Typical Marine Equipment (Btu/hr)

Equipment Max. qsd Max. qld

Clothes Dryer, Electric * 4400 -

Clothes Washer 1300 -

Coffee Maker 6500 2000

Coffee Warmer 230 70

Combination Steam Cooker/Kettle, Electric 33500 17500

Convection Oven, Electric 13500 6900

Dishwasher 1120 580

Fry Kettle, Electric 12500 6900

Garbage Disposal 1120 -

Griddle, Electric 13500 6900

Ice Cuber 1300 -

Iron, Electric 3400 -

Table 3-5 Continued

Equipment Max. qsd Max. qld

Meat Slicer, Electric 1120 580

Mixer, Electric 1120 580

Oven, Electric 13500 6900

Range, Electric 24700 12700

Refrigerator (up to 20 cu.ft.) 1300 -

Toaster 2230 1970

* Vented to Atmosphere

All estimated equipment heat gains must be revised after the equipment has been selected and the actual heat dissipation data is known. This is especially important for spaces with a high concentration of equipment or for new or rapidly changing equipment such as computers.

When the heat dissipation data is known or estimated, the equipment heat gain shall be calculated using the following equations:

...., FHFUqq dss ××= (3-6)

.., FUqq dll ×= (3-7)

Where an exhaust hood is fitted over the equipment, use a Hood Factor, H.F., of 0.5. The latent load for such equipment is considered as zero. Use H.F. = 1.0 in Equation 3-6 when an exhaust hood is not fitted over the equipment. See Table 3-6 for recommended Use Factors, U.F.

For electric motors, the sensible heat gain may be calculated using one of the following equations:

motor s

FUbhp q η ..2545 ××= (3-8)

OR

..3412 FUkWq motors ××= (3-9)

Where bhp is the motor brake horsepower and kWmotor is the rated kW of the motor. Motor efficiencies, ηmotor, are listed in Table 3-7.

Table 3-6: Recommended Use Factor for Typical Equipment *

Space Use Factor

Cargo Gear Equipment Room 0.5

Computer Room 1.0

Engine Control Room 1.0

Fan Rooms 1.0

Galley 0.5

Gyro Equipment Room 0.5

Offices 0.3

Pantry 0.5

Radar Equipment Room 0.5

Radio Room 0.3

All Other Spaces 0.3

* To be used when exact information is not available.

Table 3-7: Electric Motor Efficiencies * Motor Size (hp) ηmotor

1/8 or less 0.50

1/6 to 1/3 0.65

1/2 to 3/4 0.73

Larger than 3/4 to 2 0.82

Larger than 2 to 10 0.87

Larger than 10 0.91

* To be used when exact information is not available.

3.4.2 Direct Return from Electronics / Electrical Equipment

Direct return from electronics equipment and racks with internal ventilation blowers shall be considered on a case by case basis. Direct return entails providing return ducting located within six (6) inches of the equipment ventilation discharge opening. The benefit of providing direct return is a reduction in room load and in the amount of air conditioning airflow to the space. The direct return ducting shall be sized for 125% of the equipment’s internal blower rated airflow. For equipment located in an air conditioned space, 75% of the equipment’s heat dissipation can be considered as cooling coil load and the remaining 25% as room load. It is noted that direct return does not reduce the overall air conditioning system cooling load.

For equipment located in a ventilated space and provided with a dedicated exhaust terminal, 75% of the equipment heat dissipation shall be considered to be captured by the exhaust terminal and discharged to the weather. The remaining 25% of the equipment heat dissipation shall be considered as room load.

3.4.3 Sample Calculations

CASE 19 Stateroom with 5 ft under-counter refrigerator.

qs,d = 1300 Btu/hr

(Table 3-5)

U.F. = 0.3 (Table 3-6)

H.F. = 1.0 (Sec. 3.4.1) qs = 1300 × 0.3 × 1.0 = 390 Btu/hr (Equation 3-6)

CASE 20 Laundry with electric clothes dryer, clothes washer, and iron.

Dryer qs,d = 4400 Btu/hr (Table 3-5)

Washer qs,d = 1300 Btu/hr (Table 3-5)

Iron qs,d = 3400 Btu/hr (Table 3-5)

All U.F. = 0.3 (Table 3-6)

All H.F. = 1.0 (Sec. 3.4.1)

Dryer qs = 4400 × 0.3 × 1.0 = 1320 Btu/hr (Equation 3-6)

Washer qs = 1300 × 0.3 × 1.0 = 390 Btu/hr (Equation 3-6)

Iron qs = 3400 × 0.3 × 1.0 = 1020 Btu/hr (Equation 3-6)

Total qs = 1320 + 390 + 1020 = 2730 Btu/hr

CASE 21 Ventilated Galley – electric range with hood.

qs,d = 24700 Btu/hr

U.F. = 0.5 (Table 3-6)

H.F. = 0.5 (Sec. 3.4.1) qs = 24700 × 0.5 × 0.5 = 6175 Btu/hr (Equation 3-6)

CASE 22 Air conditioned galley – convection oven, electric without hood.

qs,d = 13500 Btu/hr ql,d = 6900 Btu/hr

H.F. = 1.0 (Sec. 3.4.1) qs = 13500 × 0.5 × 1.0 = 6750 Btu/hr (Equation 3-6) ql = 6900 × 0.5 = 3450 Btu/hr (Equation 3-7)

CASE 23 Air conditioned Galley – convection oven, electric with hood.

qs,d = 13500 Btu/hr

H.F. = 0.5 (Sec. 3.4.1) qs = 13500 × 0.5 × 0.5 = 3375 Btu/hr (Equation 3-6) ql = 0 Btu/hr (Sec. 3.4.1)

CASE 24 Fan Room with 5.0 HP motor.

bhp = 5.0 HP

(Given)

U.F. = 1.0 (Table 3-6) ηmotor = 0.87 (Table 3-7)

87.0

0.125450.5 ××= q = 14626 Btu/hr (Equation 3-8)

CASE 25 Fan Room with 5.0 kW motor.

kWmotor = 5.0 kW

U.F. = 1.0 (Table 3-6) qs = 5.0 × 3412 × 1.0 = 17060 Btu/hr (Equation 3-9)

3.5 Personnel Load

Personnel load is the sensible and latent gain generated by a room’s occupants.

3.5.1 Personnel Heat Gain

Personnel heat gain is calculated using the equations:

Pqq pss ×= , (3-10)

Pqq pll ×= , (3-11)

Personnel heat dissipation varies with room temperature and level of personnel activity. Recommended values are shown in Table 3-8. For messing facilities, add 30 Btu/hr sensible and latent per person eating as an allowance for heat dissipation from food.

Occupancy for staterooms shall be based on the number of sleeping accommodations. For offices, lounges, messing facilities, work rooms and similar spaces, occupancy shall be 3/4 of the seating capacity, rounded off to the nearest whole number.

Table 3-8: Room Load Components

Room D.B.

(°F)

Mess Attendants

& Working Spaces All Others

Sensible Latent Sensible Latent

75 360 440 300 300

76 345 455 290 310

77 330 470 275 325

78 315 485 265 335

79 300 500 250 350

80 285 515 240 360

81 270 530 230 370

82 255 545 215 385

83 240 560 205 395

84 225 575 190 410

85 210 590 180 420

3.5.2 Sample Calculations

CASE 26 Stateroom with single berth.

qs,p = 265 Btu/hr/person

(Table 3-8) ql,p = 335 Btu/hr/person

P = 1 person qs = 265 × 1 = 265 Btu/hr (Equation 3-10) ql = 335 × 1 = 335 Btu/hr (Equation 3-11)

CASE 27 Officer Mess Room – four tables, each with four seats, and one mess attendant.

Diners qs,d = 265 + 30 = Btu/hr/person (Table 3-8)

Diners ql,d = 335 + 30 = Btu/hr/person (Table 3-8)

Messman qs,d = 315 Btu/hr (Table 3-8)

Messman ql,d = 485 Btu/hr (Table 3-8)

Diners P = 3/4 × (4 × 4) = 12 people (Sec. 3.5.1)

Total qs = (295 × 12) + (315 × 1) = 3855 Btu/hr (Equation 3-10)

Total ql = (365 × 12) + (485 × 1) = 4865 Btu/hr (Equation 3-11)

CASE 28 Office with eight seats.

qs,p = 265 Btu/hr/person ql,p = 335 Btu/hr/person

P = 3/4 × 8 = 6 people (Sec. 3.5.1) qs = 265 × 6 = 1590 Btu/hr (Equation 3-10) ql = 335 × 6 = 2010 Btu/hr (Equation 3-11)

3.6 HVAC Conversion Factor

In general, sensible heat loads or gains can be calculated using the equation:

TQFqs ∆××= (3-12)

Equation 3-12 can be utilized to calculate room loads, heating loads, preheat loads, infiltration loads, etc.

Depending on the application, the value of the HVAC conversion factor, F, varies. In general, it is defined by the equation:

cWc F wpap )]([60 ,, ×+× = (3-13)

In the past, regardless of season or application, a conversion factor of 1.08 was used in all heat load calculations.

This factor of 1.08 is an accepted value for marine air conditioning applications, but its use in ventilation, preheating and infiltration calculations does not produce accurate results. To obtain accurate heat load calculations, Equation

3-13 must be evaluated with the proper air and water vapor properties for each application.

The specific heat of dry air, cp,a, and specific heat of water vapor, cp,w, vary with temperature and pressure.

However, this variation is small and will have negligible effect on the value of the F. As such, the specific heats of dry air and water vapor noted in Section 2.4 (cp,a = 0.24 Btu/lb/°F and cp,w = 0.45 Btu/lb/°F, respectively) can be assumed constant for all HVAC applications.

However, specific volume of air, v, and humidity ratio, W, vary widely under differing conditions. The HVAC engineer shall use the appropriate values of v and W as follows:

• For air conditioning applications only, the use of 1.08 is acceptable.

• For ventilation and infiltration calculations in the cooling season, the conversion factor shall be revised to reflect v = 14.445 ft per lb. of air and W = 0.0208 lb. water per lb. dry air at 95/82°F DB/WB. In this case, Equation 3-13 yields F = 1.04. For outside air design temperatures other than 95/82°F DB/WB, the HVAC engineer shall be responsible for revising the factor to reflect the correct specific volume of air at the prescribed outside air design temperatures.

• For preheating calculations, the conversion factor shall be revised to reflect the specific volume and humidity ratio of air at the preheat temperature. Typical preheat temperatures are between 42°F (v = 12.66 ft /lb. of air, W = 0.0008 lb. water per lb. dry air) and 50°F (v = 12.86 ft /lb. of air, W = 0.0008 lb. water per lb. dry air), from which the corresponding range of conversion factors, by Equation 3-13, is 1.14 ≥ F ≥ 1.12. Use of F = 1.14 is conservative, and is recommended as a general guideline for calculating the preheat load.

• For infiltration at 0/0°F DB/WB, v = 11.599 ft per lb. of air and W = 0.0008 lb. water per lb. dry air. In this case, Equation 3-13 yields F = 1.24. If the outside air design conditions are other than 0/0°F DB/WB, the

HVAC engineer shall revise the factor accordingly.

Table 3-9 indicates properties of air at various cooling season outside air conditions and indicates the corresponding

HVAC conversion factor. Table 3-10 indicates the same for different preheating temperatures and heater entering conditions.

Table 3-9: HVAC Conversion Factors for Ventilation, Infiltration, and A/C Applications

Outside Design

Temperatures

(°F DB/WB)

Specific

Volume, v

(ft /lb. of air)

Conversion

Factor, F

Humidity

Ratio,W

(lb. water vapor/lb. dry air) Application

105.0 / 87.5 14.790 1.02 0.0245 Ventilation

95.0 / 82.0 14.445 1.04 0.0208 Ventilation

90.0 / 81.0 14.318 1.05 0.0210 Ventilation

0 / 0 11.599 1.24 0.0008 Infiltration

10 / 10 11.861 1.22 0.0013 Infiltration

For A/C applications: F = 1.08

Table 3-10: Preheating HVAC Conversion Factors for Various Preheat Conditions

Preheat Design

Temperature

(°F DB)

Outside Design

Conditions

(°F DB/°F WB )

Specific

Volume, v

(ft /lb. of air)

Humidity

Ratio,W

(lb. water vapor/lb. dry air)

Conversion

Factor, F

42 0 / 0 12.669 0.0008 1.14

50 10 / 10 12.871 0.0013 1.12

45 0 / 0 12.659 0.0008 1.14

50 10 / 10 12.853 0.0013 1.12

3.7 Infiltration Load

Infiltration is leakage of outdoor air into an air conditioned space. This leakage imposes both a sensible and a latent heat load. Normally, infiltration is only considered for spaces which are not pressurized and have direct access to the weather. It is recommended that an infiltration load only be applied to an unpressurized wheelhouse because it is common practice to have wheelhouse doors open when navigating a ship near land. All other outside doors are considered as closed. Special infiltration requirements should be detailed in the construction specifications.

The sensible heat component of infiltration is included in the cooling and heating loads. The latent heat component is only included in the A/C cooling loads. The latent heat gain is not considered for heating, except when humidification is required by the specifications.

3.7.1 Infiltration Heat Gain (Loss)

Infiltration heat gains or losses are calculated using the following equations:

Total Heat hhQ q ro )(60 −××

= (3-14)

Sensible Heat

Cooling Season: TQqs ∆××= 04.1 (3-15)

Heating Season: TQqs ∆××= 24.1 (3-15a)

Latent Heat

Cooling Season: sl qqq −= (3-16)

Where ΔT is the difference between design outdoor dry bulb temperature and design room dry bulb temperature, and the constants 1.04 and 1.24 are the appropriate HVAC conversion factors from Table 3-9.

For spaces that are not pressurized, the infiltration rate of change shall be 60 minutes for cooling calculations and 10 minutes for heating calculations. The dimensions used in calculating the room volume are the finished deck heights and horizontal dimensions inside joiner linings and bulkheads.

If the space is pressurized, the infiltration rate of change shall be 120 minutes for cooling calculations and 20 minutes for heating calculations.

3.7.2 Sample Calculations

CASE 29 Wheelhouse – unpressurized, finished dimensions 8'-2" x 30'-4" x 20'-7". Cooling season: outside air conditions 95°F DB, 82°F WB; inside air conditions 78°F DB, 66.5°F WB. Heating season: outside air temperature 0°F DB; inside air temperature 70°F DB.

Cooling Calculations

V = 8' × 30' × 21' = 5040 ft

(Sec. 3.7.1)

Q = 5040 / 60 = 84 CFM (Sec. 3.7.1)

ΔT = 95 – 78 = 17°F (Sec. 2.1.1) v = 14.445 ft /lb. of air (Psychrometric Chart) ho = 45.75 Btu/lb (Psychrometric Chart) ho = 31.11 Btu/lb (Psychrometric Chart) q = 445.14

)11.3175.45(8460 −×× = 5108 Btu/hr (Equation 3-14) qs = 1.04 × 84 × 17 = 1485 Btu/hr (Equation 3-15) qs = 5108 – 1485 = 3623 Btu/hr (Equation 3-16)

Heating Calculations

Q = 5040 / 10 = 504 CFM (Sec. 3.7.1)

ΔT = 70 – 0 = 70°F (Sec. 2.1.2) qs = 1.24 × 504 × 70 = 43747 Btu/hr (Equation 3-15a)

CASE 30 Wheelhouse – pressurized, finished dimensions 8'-2" x 30'-4" x 20'-7".