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About the Authors

Qingyan (Yan) Chenis a professor of mechanical engineering at Ray W. HerrickLaboratories, Purdue University, West Lafayette, Indiana. He received his B.Sc. degreefrom Tsinghua University and M.Sc. and Ph.D. degrees from Delft University of Tech-nology. He has published over 80 archival journal papers and more than 60 conferencepapers. Since 1995, he has been the principal investigator or co-principal investigatorof 30 sponsored research projects, including five from ASHRAE. He has been electedto the International Academy of Indoor Air Sciences. Currently, Prof. Chen serves as

an associate editor for theInternational Journal of HVAC&R Researchand as an edito-rial board member for the International Journal of Ventilationand the InternationalJournal on Architectural Science.

Leon R. Glicksmanis a professor of building technology in the Department ofArchitecture as well as professor of mechanical engineering at Massachusetts Instituteof Technology (MIT). He received his B.Sc. and Ph.D. degrees from MIT and his M.Sc.degree from Stanford. Currently, Prof. Glicksman is leading an MIT effort to developenergy-efficient, sustainable building technologies and compatible designs. He hasconducted research sponsored by the EPA, NSF, DOE, ABB, and numerous industrialsponsors. He has written over 180 technical articles and chapters in four books.Currently, Prof. Glicksman serves as an associate editor for ASHRAEsInternationalJournal of HVAC&R Research.

This publication was prepared under ASHRAE Research Project 949 in cooperationwith TC 5.3, Room Air Distribution, and TC 4.10, Indoor Environment Modeling.


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Qingyan Chen

Leon Glicksman

American Society of Heating, Refrigerating

and Air-Conditioning Engineers, Inc.

System Performance Evaluation

and Design Guidelines for

Displacement Ventilation


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ISBN 1-931862-42-7

2003 American Society of Heating, Refrigerating

and Air-Conditioning Engineers, Inc.

1791 Tullie Circle, N.E.Atlanta, GA 30329

www.ashrae.org

All rights reserved.

Printed in the United States of America

Cover design by Tracy Becker.

ASHRAE has compiled this publication with care, but ASHRAE has not inves-

tigated, and ASHRAE expressly disclaims any duty to investigate, any product, service,

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in this publication. The entire risk of the use of any information in this publication is

assumed by the user.

No part of this book may be reproduced without permission in writing from

ASHRAE, except by a reviewer who may quote brief passages or reproduce illustra-

tions in a review with appropriate credit; nor may any part of this book be reproduced,

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photocopying, recording, or otherwithout permission in writing from ASHRAE.

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v

Contents

PREFACE . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . vii

ACKNOWLEDGMENTS. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ix

CHAPTER 1INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1

1.1 Displacement Ventilation. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2

1.2 Special Features in U.S. Buildings . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3

1.3 Objective of This Book . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4

CHAPTER 2LITERATURE REVIEW . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7

2.1 Temperature Distribution . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7

2.2 Flow Distribution . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .14

2.3 Contaminant Distribution . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21

2.4 Comfort Aspects . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26

2.5 Energy and Cost Analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 29

2.6 Design Guidelines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .33

CHAPTER 3EXPERIMENTAL STUDY ANDVALIDATION OF CFD PROGRAM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35

3.1 Experimental Facility. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .36

3.2 Test Procedure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .39

3.3 Experimental Results . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .42

3.4 Computational Fluid Dynamics Model . . . . . . . . . . . . . . . . . . . . . . . .42

3.5 Validation of CFD Program. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .45

3.6 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .52

CHAPTER 4MODELS FOR PREDICTION OF TEMPERATURE

DIFFERENCE AND VENTILATION EFFECTIVENESS . . . . . . . . . . . . . . . . .55

4.1 A Database of Displacement Ventilation. . . . . . . . . . . . . . . . . . . . . . .56

4.2 Model of the Air Temperature Difference

Between the Head and Foot Level . . . . . . . . . . . . . . . . . . . . . . . . . . . .704.3 Ventilation Effectiveness Model . . . . . . . . . . . . . . . . . . . . . . . . . . . . .77

4.4 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .79


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System Performance Evaluation and Design Guidelines for Displacement Ventilation

vi

CHAPTER 5PERFORMANCES EVALUATION

OF DISPLACEMENT VENTILATION. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 81

5.1 Evaluation Criteria . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 815.2 Performance Evaluation of Displacement Ventilation . . . . . . . . . . . . 84

5.3 Discussion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 91

5.4 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 94

CHAPTER 6ENERGY AND COST ANALYSIS . . . . . . . . . . . . . . . . . . . . . 95

6.1 Load Calculations. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 95

6.2 Secondary Systems and Plants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 100

6.3 Energy Analysis for U.S. Conditions. . . . . . . . . . . . . . . . . . . . . . . . 100

6.4 First Cost Analysis for U.S. Conditions . . . . . . . . . . . . . . . . . . . . . . 106

6.5 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 107

CHAPTER 7DESIGN GUIDELINES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 111

CHAPTER 8CONCLUSIONS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 117NOMENCLATURE . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 119

REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 121

INDEX. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 127


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Preface

This book presents system performance evaluation and design guidelines for

displacement ventilation.The authors first reviewed the literature concerning the performance of tradi-

tional displacement ventilation. Since U.S. buildings have different layouts andlarger internal heat gains than those studied in the literature, it was necessary todevelop design guidelines for displacement ventilation for U.S. buildings underdifferent climatic conditions.

The design guidelines present two important models that were not available inthe literature: a model to calculate the temperature difference between the head andfoot level of an occupant and a model to determine the ventilation effectiveness atthe breathing level. The investigation developed the models from the results of 56cases of displacement ventilation obtained by a computational fluid dynamics(CFD) program. Those cases include a wide range of thermal and flow conditionssimilar to those found in U.S. offices, classrooms, and workshops. The CFDprogram was validated by six sets of detailed experimental data obtained from a full-scale environmental chamber simulating a small office, a quarter of a large officewith partition, and a quarter of a classroom. The data include airflow patterns anddistribution of air velocity, temperature, contaminant concentration, and turbulence.The validation also used some data obtained from the literature. The CFD programwas also used to assess the performance of displacement ventilation, such as airflowpattern and distributions of air temperature, percentage dissatisfied due to draft,predicted percentage dissatisfied, contaminant concentration, mean age of air, andventilation effectiveness. The investigation also conducted energy and first costsanalysis.

The results show that a displacement ventilation system can provide a thermallycomfortable indoor environment at a high cooling load through careful design. Theindoor air quality in a space with displacement ventilation is better if the contami-nant sources are associated with the heat sources. The displacement ventilation

system can also save energy but requires a separate heating system if it is applied tobuilding perimeter zones. This book presents a ten-step design guideline to designthe displacement ventilation system for U.S. buildings.


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ix

Acknowledgments

This book is based on the research performed for ASHRAE Research Project

RP-949, Performance Evaluation and Development for Design Guidelines forDisplacement Ventilation. The research was sponsored by TC 5.3, Room Air

Distribution, and TC 4.10, Indoor Environment Modeling. Throughout the research,

the project monitoring committee and the members of the two technical committees

made a substantial contribution to the project, including numerous suggestions in the

project meetings, critical comments on the final report, and a site visit. The authors

are very grateful for their support and help. The authors would also like to thank their

former research associates and students, Dr. Xiaoxiong (John) Yuan, Mr. Shiping

Hu, Ms. Yongqing Hu, and Prof. Xudong Yang, for their hard work on the project.

Without their contributions, such a book would not exist. Last, but not least, the

authors are grateful to the ASHRAE Special Publications staff for their careful and

beautiful work on the book layout and edit.


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1

CHAPTER 1

Introduction

Since the energy crisis in the 1970s, the insulation of buildings has been

improved in order to reduce heat loss in winter, heat gain in summer, and the infil-tration of outdoor air. As a consequence, the heat extracted from or supplied to aroom for maintaining a comfortable air temperature is reduced and the ventilationrate is also reduced by a corresponding amount, sometimes much more if the build-ing envelope is made tighter. However, such a reduction of air supply causes anincrease in the concentration of indoor pollutants and sometimes generates a non-uniform distribution of air temperature and contaminant concentration. Draft (ther-mal comfort problems) and sick building syndrome (indoor air quality problems)are very familiar ailments today that are the direct results of the poor distribution ofairflow, temperature, and contaminant concentrations. Solving these thermalcomfort and indoor air quality (IAQ) problems without consuming too much energyis a challenge for both ventilation engineers and architects.

Currently, the United States consumes more than one-third of its energy inbuildings, and there is a possibility of saving up to 20% of this energy. Saving energymay result in the reduction of the fresh air supply. This may cause poorer IAQ. Since

people spend up to 90% of their time indoors, IAQ is increasingly recognized as anessential factor for the prevention of human diseases and the promotion of people'scomfort and welfare. In the United States, about 800,000 to 1,200,000 commercial

buildings with 30 to 70 million people have problems related to IAQ (Woods 1989).The problems include eye, nose, and throat irritation, headache, recurrent fatigue,drowsiness or dizziness, and reduced powers of concentration (Spengler 1995).Dissatisfaction with the working environment could result in reduced productivityand economic loss. A survey conducted in the New England area of 94 state govern-ment office buildings showed an average productivity loss of 3%, which is attributedto poor IAQ (Axelrad 1989). Fisk (2000) estimated that the economic impact related

to respiratory illness, allergies and asthma, and sick building syndrome is $20 to$200 billion. Therefore, it is necessary to provide a good ventilation system that canprovide good IAQ and save energy.


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2

1.1 DISPLACEMENT VENTILATION

Displacement ventilation has been used quite commonly in Scandinavia duringthe past twenty years. It was first applied to the welding industry in 1978 (Belin

1978) and has since been increasingly used as a means of ventilation in industrial

facilities to provide good indoor air quality and save energy. More recently, its use

has been extended to ventilation in offices and other commercial spaces where, in

addition to air quality, comfort is an important consideration. In 1989 in Nordic

countries, it was estimated that displacement ventilation accounted for a 50% market

share in industrial applications and 25% in office applications (Svensson 1989).

Displacement ventilation system can be divided into the following three types:

Traditional displacement ventilation, as shown in Figure 1.1

Displacement ventilation with a chilled ceiling panel

Displacement ventilation with a raised floor

This book focuses on the first type: traditional displacement ventilation.

A typical displacement ventilation system for cooling, as shown in Figure 1.1,

supplies conditioned air from a low sidewall diffuser. The supply air temperature is

slightly lower than the desired room air temperature, and the supply air velocity is

low (lower than 100 fpm or 0.5 m/s). Through the diffuser, the conditioned air is

directly introduced to the occupied zone, where the occupants stay. Exhausts are

located at or close to the ceiling through which the warm room air is exhausted from

the room. Because it is cooler than the room air, the supply air is spread over the floor

and then rises as it is heated by the heat sources in the occupied zone. These heat

Figure 1.1 Sketch of displacement ventilation.


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Introduction

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sources (e.g., persons and computers) create upward convective flows in the form of

thermal plumes. These plumes remove heat and contaminants that are less dense

than air from the surrounding occupied zone.Traditionally, the amount of supply air in a displacement ventilation system has

been less than that of mixing-type systems. This necessitates careful design of thesystem configuration and operation to adequately handle the space cooling loads.The supply temperature, velocity, and vertical temperature gradient in the occupiedzone are all very important comfort-related design parameters. Compliance with thespecification of ASHRAE Standard 55-1992 (ASHRAE 1992) for acceptable verti-cal temperature difference in the occupied zone places limitations on the magnitudeof supply-room temperature difference and/or space cooling loads for a given supplyairflow rate. This is especially important when the system is applied to a U.S. build-ing in which the cooling load can be high and weather can be hot.

Previous research (Svensson 1989; Sandberg and Blomqvist 1989; Wyon and

Sandberg 1990) has indicated that in office environments with normal room heightsof around 9 ft (2.7 m), displacement ventilation cannot maintain acceptable comfortfor cooling loads above 8 to 10 Btu/(hft2) (25 to 30 W/m2) unless the air supplyvolume is increased or additional heat removal capacity is provided through the useof cooled ceiling panels. With higher ceiling heights, displacement ventilationsystems are capable of removing larger heat loads.

A stable, vertically stratified temperature field is essential for this type of systemto function properly. Numerous studies show that, when properly designed,displacement ventilation can take advantage of the naturally occurring thermal strat-ification in the room and, thus, can increase the ventilation efficiency.

1.2 SPECIAL FEATURES IN U.S. BUILDINGS

Research on displacement ventilation has been mainly conducted in Scandina-vian countries. Recently, REHVA (2002) published a guidebook on designingdisplacement ventilation in non-industrial premises. Many U.S. cities have highertemperatures in summer than those in Scandinavian cities, and U.S. offices may havemore lighting and equipment that produces more heat. Therefore, the cooling loadcould be higher in the U.S. than in Scandinavian countries (Chen et al. 1999). Inmany U.S. offices, there are large core spaces that are completely isolated from theexternal climate. Cooling is always needed in the core spaces, and there is great

potential for the use of displacement ventilation in such spaces.

On the other hand, heating and cooling are required in the perimeter spaces. InScandinavian countries, a radiator is often used to offset heating load in winter andfresh air is supplied by the displacement ventilation system. This implies that thesupply air temperature in winter can still be somewhat lower than the room air

temperature, and a stratified flow can be maintained. However, in many U.S. officebuildings, air-conditioning systems are often used for both heating and cooling andthere is no radiator available. If a displacement ventilation system is used in the


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perimeter spaces, a separate heating system is needed to maintain the flow pattern.Convectors, baseboard heaters, radiant panels, or resistance wires can be used.

However, the first costs and operating costs with two systems would be different.Displacement diffusers can be used for heating as well, but the airflow pattern will

be of the mixing type.In addition, many U.S. offices and restaurants are large spaces with many parti-

tions to form individual work stations or dinning areas, while most European officesand restaurants are small spaces. Therefore, direct application of the Scandinavianresults for U.S. design is not feasible.

1.3 OBJECTIVE OF THIS BOOK

Displacement ventilation may improve indoor air quality and has the potentialto save energy. However, the performance of displacement ventilation is still nottotally understandable, and the special features of U.S. buildings have not beenconsidered in previous research. The objective of this book is to answer the follow-ing two questions:

Is displacement ventilation suitable for U.S. buildings? How should displacement ventilation systems be designed?

To evaluate whether a ventilation system is suitable for U.S. buildings, we needto consider, simultaneously, its impact on indoor air quality, comfort, energyconsumption, and costs. In order to design such a ventilation system, it is necessaryto provide a design guide. This book tries to answer the above questions by providingthe following information for displacement ventilation systems:

1. Literature review to identify the existing results and problems

2. Experimental study to get reliable data, including the distribution of velocity,

turbulence intensity, temperature, tracer-gas concentration, etc.3. Validation of a computational fluid dynamics (CFD) program by the experimental

data to determine the accuracy of the program

4. Numerical simulation of a large number of cases by the CFD program to establisha database on the performance of displacement ventilation

5. Model development to develop models needed for design guidelines

6. Energy and cost analysis to assess the impact of energy and first costs

7. Guidelines to help designers in the U.S. to design displacement ventilation

Chapter 2 of this book presents a state-of-the-art review on displacement venti-lation. Chapter 3 describes the experimental study in a full-size test room simulatinga small office, a large office with partitions, and a classroom. The experimental

results are used to validate a CFD program. Chapter 4 describes a database ofdisplacement ventilation by CFD computations of numerous cases for different ther-mal and flow conditions for different types of U.S. buildings. Based on the computed


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Introduction

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results, two models are developed for prediction of the air temperature differencebetween head and foot level and the ventilation effectiveness at head level. Chapter

4 also introduces a simplified CFD program for calculating indoor airflow. Chapter5 discusses the performance of displacement ventilation. Chapter 6 describes energyand cost analysis. Chapter 7 outlines a ten-step design guideline for displacementventilation. Chapter 8 offers important conclusions about displacement ventilation.


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CHAPTER 2

Literature Review

This chapter discusses some of the literature on displacement ventilation,

including studies on the distribution of temperature, velocity, and contaminants,comfort, energy, cost analysis, and design guidelines.

2.1 TEMPERATURE DISTRIBUTION

Since displacement ventilation systems supply cold fresh air directly to theoccupied zone, potential draft exists at floor level. In addition, the large temperaturestratification that exists in a space with displacement ventilation may also causediscomfort. Therefore, a designer needs information on the air temperature distri-

bution in spaces with displacement ventilation.

Dimensionless Temperature and Vertical Temperature Gradient

Researchers found that the air temperature in a space appears to vary linearlywith space height in the stratified zone and is nearly constant in horizontal directionsexcept in the region near the supply diffusers. Figure 2.1 presents a typical temper-ature profile assumed for a room with displacement ventilation, where Ts, Tf, and Teare air temperatures at supply outlets near the floor in the occupied zone and atexhausts. This linear profile is widely used by most investigators, e.g., Mathisen andSkaret (1983), Flatheim (1984), Sandberg (1985), Holmberg et al. (1987), Nielsen(1993), and Skistad (1994).

Figure 2.2 plots the vertical dimensionless temperature profiles, = (T Ts/Te Ts), in normal offices obtained from several different investigations. The dimen-sionless temperature near the floor, f, or the ratio of the temperature difference

between the supply air and the air near the floor to the difference between the supplyand extract air, varies from 0.2 to 0.7. In addition, the air temperature gradient isnot the same for different investigations. The temperature does not vary linearly


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from the floor to the exhaust for most cases. The discrepancies among the profilesin Figure 2.2 could be due to different thermal and flow conditions, such as:

ventilation rate

cooling load

heat source type and position

wall temperature and wall radiative characteristics

space height

diffuser type

Figure 2.1 Simplified vertical temperature profile in a room with

displacement ventilation.

Figure 2.2 Temperature profiles in office rooms.


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Since the temperature difference between head and feet is a critical criterion ofthermal comfort, it is desirable to have a model to predict the vertical temperature

gradient in the occupied zone.

Impact of Ventilation Rate and Cooling Load

Sandberg (1985) measured vertical temperature profiles in a test room (14 ft 12 ft 8.3 ft or 4.2 m 3.6 m 2.5 m) with displacement ventilation. The resultsshow that f is between 0.56 and 0.48 when the air change rate is between 2 and 4ach. The results of Chen et al. (1988) for a test chamber of 18.7 ft 10 ft 10.7 ft(5.6 m 3.0 m 3.2 m) confirmed that f is a function of ventilation flow rate. Thef decreases from 0.4 to 0.2 when the air change rate increases from 3 to 5 ach. Simi-lar results can be found from Mundt (1990) and Li et al. (1992).

Mundt (1990) developed a formula to calculate fbased on the assumption thatthe convective heat transfer from the floor to air raises the air temperature from Tsto T

f. The radiative heat transfer from the ceiling to the floor keeps the energy balance

on the floor surface. The f is a function of ventilation rate given as

(2.1)

where

V = ventilation flow rate,

= air density,

Cp = specific heat at constant pressure,

A = floor area,

r = radiative heat transfer coefficient from ceiling to floor, and

cf = convective heat transfer coefficient from the floor to room air.

As shown in Figure 2.3, Equation 2.1 is in good agreement with most measureddata in the literature (the same references cited in Figure 2.2) when r= 0.9 Btu/(hft2F) (5 W/m2K) andcf= 0.7 Btu/(hft

2F) (4 W/m2K). Point 1 in Figure 2.3is relatively small because the walls were covered with aluminum in the experimentand the radiative heat transfer to the floor was small. Point 2 is relatively large

because the cooling load was small and the total temperature difference was only5F (2.8C) in the experiment. Equation 2.1 accounts for the impact of cooling loadon fbecause ventilation rate and cooling load are interrelated.

If the temperature varies linearly with the elevation, the temperature gradient,s, can be estimated as

s = (1 f) (Te Ts) / H (2.2)

or

f1

VCpA

--------------1

r-----

1cf-------+

1+

-------------------------------------------------- ,=


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(2.3)

Unfortunately, Equation 2.3 does not correlate the temperature gradient, asshown by Figure 2.4, where the points represent the measured data (the same refer-ences cited in Figure 2.2).

Impact of Heat Source Type and Position and Wall Characteristics

Nielsen et al. (1988) showed that, in a room with constant load from a concen-trated heat source, the gradient of dimensionless temperature decreases slightly asthe Archimedes number (ghTe/us

2) increases. The gradient is strongly related tothe surface temperature of the heat sources (Nielsen 1992). Nielsen (1996) later

presented a design chart, shown in Figure 2.5, from experimental results in roomswith heights of 8.3 ft to 15 ft (2.5 m to 4.5 m) to determine dimensionless temperatureof air near the floor.

It is a common practice to assume adiabatic thermal conditions for internalwalls. In many cases, the internal walls are not adiabatic. Since wall area is large,a small temperature difference between the walls and room air could result in a

significant downflow (if the walls are colder) or upflow (if the walls are warmer). Inaddition, the temperature along a vertical line of an internal wall is not a constant,and there is a temperature gradient in room air. Heat transfer occurs between room

sH

Te--------- 1 f 1

1Vcp

A------------

1r-----

1cf-------+

1+

------------------------------------------------.= =

Figure 2.3 Dimensionless temperature near the floor vs. supply flow rate.


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Literature Review

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Figure 2.4 Dimensionless temperature gradient vs. supply flow rate.

Figure 2.5 A design chart for the dimensionless temperature near the floor

(Nielsen 1996).


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air and internal walls. Most of the investigations neglect the impact of internal walltemperature on the vertical temperature gradient.

A study conducted by Jarmyr (1982) showed vertical temperature profiles atfive different times of day in a workshop. The temperature gradient increased from0.23F/ft (0.38 K/m) in the early morning to 0.42F/ft (0.7 K/m) at noon and thendecreased slightly to 0.39F/ft (0.65 K/m) in the evening. The nondimensionaltemperature near the floor,f, varied from 1/3 in the morning to 1/7 in the afternoon.It is clear that heat from external walls and windows contributes to the temperaturegradient.

Li et al. (1992) showed that heat conduction through walls and radiationbetween room surfaces, particularly between ceiling and floor, make a significantcontribution to the vertical temperature profile. For example, the dimensionless airtemperature near the floor was changed from 0.62 to 0.3 when the black walls werecovered with aluminum. To include the contribution of radiative heat transfer andconduction through walls, Li et al. (1992) suggested a four-node model and a multi-node model. Good agreement was found between the models and their measureddata.

Mundt (1996) recently extended her early model (Mundt 1990) to consider theinfluences of heat transfer through the building enclosure and the heat sources on thevertical temperature profile. The new model relates air temperatures near the floorand ceiling to room geometry and the heat transfer among the room air, heat sources,floor, and ceiling.

Figure 2.6 shows that the models of both Mundt (1996) and Li et al. (1992) couldpredict the measured data of Li et al. (1992). However, the models do not closelypredict the results of various investigators. Since both models assume the air temper-

ature gradient to be constant, they need improvements.Cooled ceiling panels with displacement ventilation are often used in spaceswith a high cooling load. The vertical temperature gradients in the spaces with the

Figure 2.6 Comparison of the temperatures between the models and the

experiment.


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Literature Review

13

cooled ceiling panels are smaller than those without the panels. The temperaturedistributions are almost uniform in the upper zones, as reported by Skistad (1994)

and Taki et al. (1996). If the panel temperature is too low, the displacement venti-lation could be transformed into mixing ventilation.

Impact of Space Height

Displacement ventilation is more suitable for high spaces, such as concert hallsand workshops (Skistad 1994). Skistad (1989) studied temperature profiles in aconcert hall with supply openings under chairs. The temperature rises rapidly fromthe supply air temperature atthe floor to the elevation where people are located.Above the people, there is only a slight temperature gradient up to the elevationwhere the lights are located. At that level, another temperature jump occurs, which

brings the air temperature up to the exhaust air temperature at the ceiling level.

Niemela and Koskela (1996) made measurements in a large industrial hall with

a height of 90 ft (27 m). Their results show that the temperature increases with eleva-tion in the zone lower than 23 ft (7 m). In the upper zone, the temperature is almosta constant. These measurements confirm again that the vertical temperature gradientis not a constant. Large spaces may be divided into a few zones for determining thetemperature distribution.

Impact of Diffuser Type

Skaret (1985) and Nielsen et al. (1988) investigated the impact of supply diffus-ers on the temperature distribution. It is better to increase the entrainment of roomair so as to decrease the temperature gradient in the occupied zone. The performanceof diffusers is critical to avoid draft near the diffusers. Recently, manufactures havedeveloped new products, such as the aspirating diffuser and the modulating diffuser.

The performance data can be found from product catalogs.

Conclusions

The air temperature near the floor and vertical temperature gradient in the occu-pied zone are two of the most important parameters to evaluate the performance ofdisplacement ventilation in terms of comfort. The ventilation rate or cooling load hasa significant impact on the two parameters. It is possible to determine the air temper-ature near the floor, but it is difficult to calculate the gradient.

The type and position of the heat sources and wall heat transfer also influencethe air temperature near the floor and the vertical temperature gradient. Some of thecontributions have been estimated in previous investigations. However, the estima-tion sometimes gives erroneous prediction of the vertical temperature gradients

because some of the influential parameters are not accounted for.If displacement ventilation is used for space with a high ceiling, it is desirableto use a more sophisticated model.


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14

A good diffuser should mix the supply air with the surrounding air quickly toreduce possible draft.

It is necessary to develop a universal but simple equation that can put all theresults together for design purpose.

2.2 FLOW DISTRIBUTION

One important feature in displacement ventilation is to properly control anddesign the airflow distribution. Proper distribution will ensure good air quality andcomfort level in the space. For example, well-designed displacement ventilation canachieve a one-dimensional displacement flow in the occupied zone and transport thecontaminants to the upper zone. Both thermal plumes and supply air from diffusers

play an important role in the airflow distribution.

Impact of Thermal Plumes

For proper design of displacement ventilation, it is important to calculate theentrained flow as a function of height. The thermal plume generated by a heatedobject will increase its volume with the height above the object, as shown in Figure2.7.

Figure 2.7 A thermal plume above a heated object.


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According to Baturin (1972), the flow rate, V, at a height,y, from a concentratedheat source in homogeneous surroundings can be expressed by

V= 0.005Qc1/3(y+y0)

5/3, (2.4)

where

y+y0 = distance from a virtual origin of the flow (y0can be approximated as 2dand

dis the hydraulic diameter of the heat source) and

Qc = convective heat emission from the source.

This equation is valid fory> 2d(Kofoed and Nielson 1990). Qccan be estimatedby Qc =k Qt, where Qtis the total energy consumption of the heat source (includingconvective and radiative). The coefficient kis 0.7 to 0.9 for pipes, 0.4 to 0.6 for small

components and 0.3 to 0.5 for large machines (Nielsen 1993).

A line source generates a two-dimensional plume, for which Skaret (1986)

suggested using the following formula to calculate the flow rate:

V= 0.014(Qc/l)1/3(y+y0) l (2.5)

where l is the length of the source andy0is one to two times the heat source width.

Stabi (1988) presented a list of volume flow rates above different heat sources

(including people, machines, windows, and radiators) in homogeneous surround-ings.

In an environment with temperature stratification, such as a space with displace-ment ventilation, the air temperatures in the plume and surrounding are identical at

a certain level. Higher than this level, no buoyancy force exists in the plume. There-fore, the thermal plume can only reach a maximum height in an environment with

temperature stratification. Mundt (1992) found that the flow rate of a thermal plumein a space with a vertical temperature gradient is a little smaller than that without thegradient. The maximum height of the plume is significantly shorter. If the height ofa plume is less than the height of the occupied zone (6 ft or 1.8 m from floor), the

contaminants within the plume will spread in the occupied zone and cannot reach theupper zone. Therefore, the maximum height of a plume is an important design

parameter. Mundt (1992) presented the following equations to calculate the flow rateand the maximum height of a plume in a space with air temperature gradient:

V= 0.00238 Qc3/4 s-5/8 (0.004 + 0.039y1+0.380y1

2 0.062y13) (2.6)

ymax= 0.98 Qc1/4 s3/8y0 (2.7)

where

y1= 2.86 (y+y0) Qc1/4 s3/8 (2.8)


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V = flow rate in a plume, m3/s

ymax = maximum height of the plume, m

s = air temperature gradient, K/m

Qc = convective heat emission, W

y = height above the heated object, m

y0 = distance between the virtual origin of the plume and the heated object, m

Many researchers investigated the plume generated by a sedentary person.Mierzwinski and Popiolek (1981) reported that the convective airflow is in a rangeof 1.1 to 2.2 ft3/s (30 to 60 l/s) through a section 2.5 ft (0.75 m) above the personshead. According to Clark and Edholm (1985), the flow rate may increase as thesurrounding temperature decreases because of the increase of a persons metabo-

lism. Danielsson (1987) showed that the vertical temperature gradient in the airsurrounding a person has a strong impact on the flow rate above the person. He also

provided a chart for determination of the flow rate above a person. According to thechart, the flow rate at 6 ft (1.8 m) above the floor may decrease from 1.56 ft3/s to 0.74ft3/s (42 l/s to 20 l/s) when the room temperature gradient is increased from 0.3 to0.9F/ft (0.5 to 1.5C/m). Fitzner (1989) confirmed the results.

Kofoed and Nielsen (1990) further reported that the flow rate may be influenced

by not only the temperature gradient but also the ventilation rate. The maximumheight of a plume generated by a person is about 6.6 to 10 ft (2 to 3 m), dependingon the vertical temperature gradient. The measured flow rates by Mundt (1992) areabout twice as much as those measured by Danielsson (1987) and Fitzner (1989).Figure 2.8 shows the flow rate of the plume calculated by Equations 2.6 and 2.7. Thecalculated results agree reasonably with the measured data (Mundt 1992; Kofoed

and Nielsen 1990).

The flow rate in a plume generated by a lamp is smaller than that generated bya person, and the flow rate in the plume created by a desk lamp is much smaller thanthat created by a computer (Mundt 1992), although the energy consumption is at thesame level. Figure 2.9 presents the measured data of Mundt (1992) for the volumeflow rate above a desk lamp, a fluorescent lamp, and a personal computer. Equations2.6 to 2.8 may be applied to fluorescent lamps and personal computers. The resultsseem logical because a point heat source has a smaller area that would entrain muchless air than a heat source with a large area.

In many cases, a heated object is placed close to a wall. Due to the Coanda effect,

a plume close to a wall can be considered as one half of the flow in a free plume witha double convective heat emission 2 Qc. The flow rate of a plume close to a corneris about one quarter of the rate in a free plume with 4 Qc. The plumes generated by

a number of sources near each other may form a large plume with a flow rate of aboutN1/3V, whereN is the number of the heat sources and Vis the flow rate in a free plume(Nielsen 1993).


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Figure 2.8 Volume flow rate around and above a person.

Figure 2.9 Volume flow rate above a personal computer and a lamp(Mundt 1992).


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Impact of Walls

Buoyancy will drive airflow up (or down) along a hot (or cold) vertical surface,such as a wall. The flow rate in the turbulent boundary layer may be calculated from(Nielsen 1993)

V= 0.0028 Tw2/5y6/5l (2.9)

where

V = flow rate in the boundary layer, m3/s;

Tw = temperature difference between room air and the wall surface, C;

y = length measured from the leading edge, m;

l = horizontal width of the surface, m.

The up or down airflow along a wall is a typical wall jet. Heiselberg (1993)presented a formula to calculate the maximum velocity in the layer. For a modest

temperature difference of a few degrees between the wall and room air, the flowalong the wall may be as large as that from several heat sources in the room such as

people or equipment.A cold downdraft from vertical cold surfaces may cause a stratified flow

with a typical wall jet profile near the floor. Heiselberg (1993) measured theprofiles and presented formulae to calculate the maximum velocity in the flowand to evaluate the draft risk. The maximum velocity near the wall and the flooris about 40 fpm (0.21 m/s) for a 5 ft (1.5 m) high cold wall with a differenceof 18F (10C) between the cold wall surface and the room air.

Impact of Diffusers

Since relatively cold air is supplied directly to the occupied zone in displace-ment ventilation, the velocity has to be well controlled to avoid draft. The velocitynear a diffuser depends on the flow rate from the diffuser, the temperature difference

between the supply and exhaustTe, and the diffuser type. Figure 2.10 shows a typi-cal velocity distribution near a diffuser (Nielsen 1993).

Skistad (1994) divided the flow near the floor into two regions: the primaryregion (where the flow is dominated by the characteristics of the diffuser) and thesecondary region (the part outside the primary region). He presented a formula tocalculate the maximum velocity in the primary region through (1) the induced ambi-ent airflow volume caused by the dynamics of the jet discharged from the diffuser,(2) the entrained airflow volume caused by the shear in the boundary layer betweenthe supply air and the ambient air, (3) the thickness of the supply air blanket, and (4)the Archimedes number Arh(ghTe/us

2). More research is needed to calculateparameters (1), (2), and (3).

In the secondary region, Nielsen (1993) presented the following equation todetermine the maximum velocity in the center plane umax,xin a distancexfrom thewall-mounted diffuser:


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umax,x= Kdr(h/x)uf (2.10)

where

h = diffuser height

uf = face velocity defined as the volume flow rate divided by the face area of the

diffuser

Kdr = a function of the Archimedes number that strongly depends on the structureof the diffuser

Nielsen (1993) also providesKdr

data as shown in Figure 2.11. The Kdr

dependson diffuser tape.

The distance from a wall-mounted diffuser to the 40 fpm (0.2 m/s) velocitycontour along the center line, ln, is an important parameter. According to ASHRAEStandard 55-1992, the air velocity should be no higher than 50 fpm (0.25 m/s). Skis-tad (1994) suggested discharging air obliquely to both sides, with some degree ofturbulence to reduce ln, and using perforated panels instead of a filter mat to reducethe draft effect. To make lnsmaller is a primary goal for diffuser manufacturers.

Normally, the manufacturers provide charts to determine lnand velocity distribution

near the diffuser in their product catalogs. Figure 2.12 shows an example.

The flow from a number of diffusers placed closed to each other on the wall willmerge to a two-dimensional flow, in which the velocity is lower than that in the radial

flow near a single diffuser (Nickel 1990). However, if the diffusers with obliquedischarges are located too close, the discharged flows meet and turn straight into theroom, and lncould be several meters (Skistad 1994).

Figure 2.10 A typical velocity distribution near a diffuser (Nielsen 1993).


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Figure 2.11 Measured value for some diffusers (Nielsen 1993).

Figure 2.12 Design chart for choosing a diffuser.


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Conclusions

Stratification height, which is a function of the flow rates of supply air and ther-mal plumes, is an important design parameter for displacement ventilation. The flowrate at a certain height in a thermal plume can be determined by the heat source type,location, and geometry. The temperature gradient in a space has an impact on theflow rate and the maximum height of a plume.

To avoid draft, velocity in the occupied zone, especially near the diffusers, needsto be well controlled. Previous studies provide sufficient information to developdesign guidelines. The design charts provided by diffuser manufacturers are alsouseful.

2.3 CONTAMINANT DISTRIBUTION

The advantage of displacement ventilation is that it may provide better indoorair quality in the occupied zone than mixing ventilation. It is therefore important tostudy the impact of different parameters, such as contaminant source type and loca-tion, human body convection, wall surface temperature, and space height, on thecontaminant distribution.

Impact of Contaminant Source Type and Location

Typically, the occupied zone with displacement ventilation has a lower contam-inant concentration level than that in the upper zone, as shown in Figure 2.13 (Heisel-

Figure 2.13 Typical profiles of the contaminant concentration vs. different

ventilation rates (Heiselberg and Sandberg 1990).


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22

berg and Sandberg 1990). Chen et al. (1988) showed that both the energy andventilation efficiencies of displacement ventilation are higher than those of mixingventilation when the contaminant source is combined with a heat source. The venti-lation efficiency increases as the ventilation rate increases or the cooling loaddecreases.

Olesen et al. (1994) reported that the concentration distribution depends on thecontaminant density (Figure 2.14). However, the amount of contaminants must besufficiently large to form the density difference. In most measurements using tracergas technique, the impact of density is negligible.

There are cases when the contaminant concentration is not lower in the occupiedzone than that in the upper zone. Figure 2.15 presents a measured concentration

profile in a room with displacement ventilation and a pollution source located at alow level and outside the thermal plume (Nielsen 1996). In this case, the lower zonehas a high contaminant concentration level.

Stymne et al. (1991) showed that the contaminant concentration level varies

significantly in both the vertical and horizontal directions, depending on the positionof pollutant sources related to the thermal plumes. As illustrated in Figure 2.16, thecontaminant concentration in the occupied zone is high when the contaminant is

Figure 2.14 Concentration profiles with different types of tracer gases

(Olesen et al. 1994).


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Figure 2.15 Concentration profile with pollutant source located at low level

and without heat source (Nielsen 1996).

Figure 2.16 Concentration contours in a room with a tracer gas emitted

above a 4 W heat source in a low level (Stymne et al. 1991).


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24

combined with a weak heat source. The thermal plumes are too weak to reach theupper zone.

Mundts (1996) measurements showed that the local air quality is good whena tracer gas source is placed above a heat source that produces a plume that can reachthe ceiling (a plume that can reach the upper zone should be able to maintain a goodair quality in the low zone). When the tracer gas source is placed outside of the ther-mal plume, the local air quality depends strongly on whether the tracer gas has a

positive or negative buoyancy on the room flow pattern. In this case, the occupiedzone might have a high contaminant concentration level. The conclusions are similar

to those of Stymne et al. (1991).

Impact of Convection from Human Bodies

Holmberg et al. (1987) found that a free convection flow around a person mayprotect the breathing zone from surrounding contaminants at the head level, but itmay also bring contaminants from the source below the breathing zone. Saeteri(1992) showed that CO2concentration in the air inhaled is lower than that at the sameelevation some distance from the person because the convection flow around thehuman body brings fresher air from the floor level directly to the breathing zone.This has been confirmed by Murakami et al. (1997) through a detailed computationalfluid dynamics simulation.

As indicated in Figure 2.17, Brohus and Nielsen (1994) showed that the concen-

tration in the inhaled air is 0.58cethe same as that at 1.7 ft (0.5 m) below the breath-ing levelalthough the concentration outside the thermal boundary layer around theperson at the breathing level is 0.96ce. They found the concentration of inhaled

Figure 2.17 Inhaled air is located below the breathing level. (The measured

concentration of the inhaled air at 1.5 m is 0.58ce, instead of

0.96ce(Brohus and Nielsen 1994).


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contaminant Cimay be expressed as a linear function of the stratification height,yst,as follows:

Ci= Cy(Cy Cf)yst /yb (yst


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found that the movement of people causes an increase of the concentration of inhaledcontaminants due to the disturbance to the free convection flow around people. This

flow transports fresh air from the floor level to the breathing zone.Fukao et al. (1996) conducted measurements in two larger offices with different

ventilation systems. The results indicated that the air quality with the floor-mounteddisplacement system is better than that with a ceiling-mounted mixing system, whilethe thermal environments are almost the same between the two systems. Tanabe andKimura (1996) measured the mean age of air in an office room with three differentventilation systems. They concluded that a wall-mounted displacement system

provides better air quality than a floor-mounted displacement system, and the floor-mounted system is better than a ceiling-mounted mixing system.

Conclusions

Contaminant concentration distribution depends on contaminant source type

and location and its associated plume strength, etc. Low contaminant concentrationmay be obtained in the occupied zone when the contaminant source is associatedwith a heat source and the thermal plume generated by the heat source is sufficientlystrong to reach the upper zone.

Because the upward free convection around a person brings the air from a lowerlevel to the breathing zone, the inhaled air is cleaner than the air at the same height.

Cold walls or cooled ceiling panels may lead to a higher contaminant concen-tration in the occupied zone because of possible downflow driven by the walls or

panels.It is more beneficial to apply displacement ventilation for spaces with high ceil-

ings, if the contaminants are buoyant gases.Prediction of contaminant distribution is more difficult than air temperature and

flow distribution.

2.4 COMFORT ASPECTS

The primary reason for using displacement ventilation is to achieve a high IAQenvironment. However, the ventilation must maintain an acceptable comfort level.Previous investigations showed that large vertical temperature gradient and draft arethe two main causes of discomfort with displacement ventilation. To reduce thetemperature gradient, the supply flow rate must be increased. This will lead to a highair velocity at the floor level and to a high draft risk. It is also not feasible to increaseventilation rate because of energy concerns.

Draft Risk Assessment

In a room with displacement ventilation, Wyon and Sandberg (1990) tested

sensitivity of 36 male and 36 female subjects to different velocity and temperatures.The percentage of discomfort is summarized in Figure 2.18. The vertical coordinateis the percentage of dissatisfied people and the horizontal coordinate is air temper-


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Figure

2.1

8Predictedpercentageofdiscomfort(a)abovechairheightand(b)below

chairheight(Wyon

and

Sandberg1990).


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ature. The results showed that the ankle and foot (below chair height) are more sensi-tive to air temperature than the rest of the body. At a velocity of 0.6 ft/s (0.2 m/s),

fewer than 20% of people will complain of local discomfort in a temperature rangeof 72F to 81F (22.1C to 27.0C). Skistad (1994) noted that the air velocities inthe range between 0.5 and 0.7 ft/s (0.15 and 0.2 m/s) are acceptable for air temper-atures of about 68F (20C), and velocities of up to 0.83 ft/s (0.25 m/s) seem accept-able for higher temperatures.

Many researchers (Chen 1988; Sandberg and Blomqvist 1989; Kegel andSchulz 1989; Olesen et al. 1994; Akimoto et al. 1995; and Taki et al. 1996) reportedthat displacement ventilation may generally provide a good thermal comfort envi-ronment in various spaces. However, the draft risk at the floor level seems rather highin spaces with displacement ventilation. Melikov and Nielsen (1989) evaluated thethermal comfort condition in 18 displacement ventilated spaces. Within the occu-

pied zone, they found that 33% of measured locations had higher than 15% of dissat-isfied people due to draft. Also, 40% of the locations were found to have a

temperature difference between head and foot of larger than 5.4F (3.0C).Some measures are available to reduce discomfort level caused by temperature

gradient. Glicksman et al. (1996) used low flow rate fans at the floor level to reducethe temperature difference between the ankle and breathing level of a seated person.The measure does not affect the flow in the upper zone in a room with displacementventilation if the vertical momentum of the fan exhaust is kept low enough.

Impact of Cooling Load and Cooled Ceiling Panel Temperature

Figure 2.19 shows the range of cooling load per floor area investigated by someresearchers. Most of the studies show that the displacement ventilation system canonly provide acceptable comfort if the corresponding cooling load is less than about13 Btu/(hft2) (40 W/m2). With higher ceiling heights, the displacement system is

capable of removing larger cooling loads (Skistad 1994).By increasing the area of the air supply outlet (e.g., supplying air through aperforated floor) or by providing additional heat removal capacity (e.g., using cooledceiling panels), displacement ventilation may be applied to a space with higher cool-ing load. Olesen et al. (1994) found that no thermal comfort problems existed underthe tested conditions with the cooling loads up to 14 Btu/(hft2) (44 W/m2) in a roomwith a perforated floor. Niu (1994) showed that displacement ventilation combinedwith cooled ceiling panels may provide a comfort environment at a cooling load upto 16 Btu/(hft2) (50 W/m2).

Taki et al. (1996) measured the vertical temperature profiles for four differentcooling loads with and without cooled ceiling panels. The results showed a signif-icant influence of the panel temperature on the air temperature distribution in theroom. The cooled ceiling panel may create downdrafts in the occupied zone. To

avoid it, the surface temperature should be higher than 59F (15C), and the ratio ofpanel area to ceiling area should be less than a certain amount. A minimum surfacetemperature is also required to avoid condensation on the panel surface.


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Conclusions

Large vertical temperature gradient and draft are the two main causes ofdiscomfort with displacement ventilation. Previous research shows that displace-ment ventilation without cooled ceiling panels is suitable for spaces with a coolingload less than 13 Btu/(hft2) (40 W/m2). However, the current study, as shown inChapter 6, indicates that the upper limit is much higher. With cooled ceiling panels,

displacement ventilation can remove a cooling load of 16 Btu/(hft2) (50 W/m2). Itis important that the surface temperature of the panels should not be lower than 59F(15C). Low flow rate fans at floor level may reduce the vertical temperature differ-ence and extend the application range.

2.5 ENERGY AND COST ANALYSIS

Annual energy consumption, first costs, and operation and maintenance costsover a life cycle are important criteria for evaluation of a ventilation system. Almostall of the energy analyses in the literature were done by numerical simulation

because it is too expensive and time consuming to conduct hour-by-hour measure-ments for a building based on a yearly basis.

Energy Analysis

Seppanen et al. (1989) evaluated the energy performance of displacement venti-lation systems and mixing ventilation systems in U.S. office buildings. The study is

Figure 2.19 Ranges of cooling load per floor area for three types of

displacement ventilation: side-wall diffuser (system 1), side-wall

diffuser with cooled ceiling panel (system 2), and rise floor

(system 3).


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for south, north, and core zones with four representative U.S. climates (Minneapolis,Seattle, Atlanta, and EI Paso). They compared different control strategies, such as

variable-air-volume system and constant-air-volume system, and systems withdifferent components, such as recirculation, economizer, and heat recovery device.The energy consumption was found to depend very much on the control strategiesand air-handling systems, as shown in Figure 2.20. The energy consumed bydisplacement systems with heat recovery and variable-air-volume flow control issimilar to that of mixing systems.

Seppanen et al. (1989) used an average cooling load of 4.4 Btu/(hft2) (14 W/m2),and the maximum load is 7.5 Btu/(hft2) (24 W/m2) in the core space of U.S. office

buildings (Table 2.1). Since the core region does not need heating, application ofdisplacement ventilation is particularly attractive. However, the load is much higherin the perimeter region. The cooling load in the perimeter seems too high to use adisplacement ventilation system. According to the results shown in Table 2.1, thetraditional displacement ventilation system can only be used in the north zone of

buildings in Seattle. For the rest, the cooling loads are much higher than the traditionaldisplacement ventilation system can handle.

Chen and Kooi (1988) pointed out the significant impact of the vertical temper-ature gradient on energy consumption in a room with displacement ventilation when

Figure 2.20 Comparison of annual energy cost of different systems to the

same costs for system 1 (VAV mixing system) in theMinneapolis climate (Seppanen et al. 1989) (Mmissing

ventilation, Ddisplacement ventilation).


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they analyzed a Dutch office with different ventilation systems. The energy

consumption of displacement ventilation can be either smaller or larger than that of

mixing ventilation, as shown in Table 2.2, depending on the control strategies and

the HVAC systems. The conclusions are similar to those of Seppanen et al. (1989),

although the approaches and weather data are different between the two investiga-

tions.

Nius (1994) calculation showed that the annual energy consumption ofdisplacement ventilation with a water-cooled ceiling system is almost the same as

that of an all-air system. His investigation used a variable-air-volume system.

Table 2.1 Heating and Cooling Loads for

Each Location and Representative Zone in the U.S.

Atlanta El Paso Minneapolis Seattle

Max. Ave. Max. Ave. Max. Ave. Max. Ave.

North zones

Heating (Btu/hft2)

Heating (W/m2)

Cooling (Btu/hft2)

Cooling (W/m2)

19.6

62.0

22.2

70.0

5.2

16.3

10.1

31.9

16.8

53.1

24.6

77.5

4.3

13.6

11.7

36.9

14.1

44.5

17.5

55.3

4.6

14.4

7.8

24.6

9.1

28.7

15.7

49.5

3.2

10.0

7.1

22.3

Core zones

Heating (Btu/hft2)

Heating (W/m2)

Cooling (Btu/hft2)

Cooling (W/m2

)

0.0

0.0

7.5

23.6

0.0

0.0

4.4

14.0

0.0

0.0

7.5

23.6

0.0

0.0

4.4

14.0

0.0

0.0

7.5

23.6

0.0

0.0

4.4

14.0

0.0

0.0

7.5

23.6

0.0

0.0

4.4

14.0South zones

Heating (Btu/hft2)

Heating (W/m2)

Cooling (Btu/hft2)

Cooling (W/m2)

19.2

60.5

37.6

118.7

5.0

15.8

14.1

44.5

16.8

53.0

41.4

130.5

4.7

14.8

17.6

55.6

16.0

50.3

36.9

116.4

5.0

15.9

12.4

39.0

9.1

28.6

35.9

113.2

3.3

10.3

11.3

35.7

Table 2.2 The Costs of Annual Energy Consumption

Ventilation

Systems Air-Handling Systems

Supply Air Temperature

F (C)

Energy Consumption

($/m2)

Displacement Variable air volume 61 (16) 108

Mixing Variable air volume 55 (12.5) 104

Mixing Variable air volume 61(16) 126

Displacement Constant air volume - 241

Mixing Constant air volume - 222


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Zhivov and Rymkevich (1998) compared the energy consumption betweendisplacement and mixing ventilation systems for a restaurant in different U.S.

climates. They found that the displacement ventilation can save up to 50% of coolingenergy but may increase heating energy.

Previous studies show that both the supply air temperature and the exhausttemperature in displacement ventilation are higher than those of mixing ventilation.The air temperature difference between the supply and the exhaust is nearly the same

between the two ventilation systems. According to Skistad (1994), the temperaturedifference for displacement ventilation can be larger for high spaces and, therefore,supply airflow rate can be reduced considerably. Note that displacement ventilationmay use more natural cooling, since the supply air temperature is 4F to 6F (2Cto 3C) higher than that of mixing-type ventilation.

First Cost Analysis

Seppanen et al. (1989) found that the first cost of a system is difficult to estimate.They compared different air-handling systems, such as a variable-air-volume systemand a constant-air-volume system, and systems with different components, such asre-circulation, economizer, and heat recovery device. Figure 2.21 shows that the firstcosts of displacement systems are substantially higher than those of mixing systemswhen cooled ceiling panels are required. Without cooled ceiling panels, the costs ofthe displacement system are similar to those of mixing system. Skistad (1994) also

Figure 2.21 Comparison of the first cost of different systems to the same costs

for system 1 (VAV mixing system) in the Minneapolis climate(Seppanen et al. 1989) (Mmixing ventilation, Ddisplacement

ventilation).


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reported that there is no significant first cost difference between the two systems,except that the cost of diffusers in the displacement ventilation is higher than that in

mixing ventilation.

Conclusions

There are not many publications in the literature concerning energy and costanalysis for displacement ventilation. Energy consumption varies significantly withclimatic regions. Compared with mixing ventilation, displacement may or may notsave energy. The energy consumption depends on the control strategies and air-handling systems.

In U.S. offices, displacement ventilation is ideal for core zones. Previous studiesshow that it may not be appropriate for perimeter zones in most regions of the U.S.

because the cooling load is too high. If cooled ceiling panels are used, the first costof displacement ventilation is much higher than that of mixing ventilation even

though the system may not be able to remove high cooling loads found in the southzone of U.S. office buildings.

Displacement ventilation combined with radiators is used in Europe for winterheating. However, the U.S. uses different heating methods, and few studies are avail-able.

2.6 DESIGN GUIDELINES

According to the analysis in the previous sections, the following parameters aremost important in design of the displacement system:

Supply airflow rate and temperature

Air temperature at floor level

Vertical temperature gradient

Maximum air velocity at floor level

Stratification height (lower zone height) or contaminant concentration gradi-ent

Energy consumption

First costs and maintenance costs

The most complete design guidelines available are those developed by Skistad(1994). He used a five-step approach:

1. Determine the required airflow rate for removal of surplus heat based on the cool-ing load and the air temperature difference between supply and exhaust openings.

2. Find the required airflow rate for removal of pollutants according to ventilationstandards.

3. Choose the larger of the two flow rates determined at Steps 1 and 2 as the venti-lation rate.


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4. Determine supply air temperature under assumptions of f= 0.5 and constantvertical temperature gradient.

5. Choose supply diffusers according to the data provided by manufacturers in orderto avoid draft.

A more recent version of the deign guide can be found from the REHVA guide-book (REHVA 2002).

Despite the simple design guidelines, there are problems. Figure 2.2 shows thatf varies from 0.2 to 0.7, and the vertical temperature gradient is not a constant. Ifthe actual Tf< 0.5Te, the vertical temperature gradient will be larger than theexpected gradient. The design guidelines assume Tf=0.5Te. If the actual Tf>0.5Te, the selected Vbased on Tf= 0.5Tewill be larger than needed.

A different approach proposed by Chen et al. (1991) used a design atlas. The

atlas, based on experimental and computational results, contains detailed informa-tion on indoor airflow, indoor air quality, and thermal comfort for various configu-rations of spaces. Unfortunately, the atlas at present does not cover a wide range ofspaces and conditions.

It seems necessary to improve current available design guidelines for thedisplacement ventilation system to ensure good indoor air quality and thermalcomfort in the space.

From the above review and a survey among architects and energy consumptionin the U.S. (Chen et al. 1999), we may conclude that design guidelines available inthe literature cannot be used with confidence. Many assumptions need further clar-ification so that designers can use the design guidelines with confidence. U.S. build-

ings, especially perimeter zones of such buildings, have a high cooling load. Designof displacement ventilation must address these zones. The special features in U.S.buildings considered in this book include offices, classrooms, and industrial work-shops.

For these types of spaces, this book will present experimental tests used toobtain reliable data on the performances of displacement ventilation. A CFD

program was validated against these data. By using the program, this book showshow to conduct numerical simulations of a large number of cases of displacementventilation in the three types of spaces and how to establish a database on the perfor-mances. Based on the database, this book further presents a model for prediction oftemperature difference between head and foot levels and a model for ventilationeffectiveness for design purpose. The book will also compare the energy and first

costs of the displacement ventilation system with a mixing ventilation system.Finally, the book presents design guidelines for the displacement ventilation systemdeveloped from the study. These will be discussed in the following chapters.


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CHAPTER 3

Experimental Study and

Validation of CFD Program

Displacement ventilation may provide better indoor air quality and save energy,

but there is a question of the usefulness of this technology in U.S. buildings withhigher cooling requirements. A first step in preparing a design guideline is carefulstudy of displacement ventilation for several typical U.S. buildings.

Two main approaches are available for the study of airflow and pollutant trans-port in buildings: experimental investigation and computer simulation. Experimen-tal investigation, although it is reliable, is very expensive and time consuming.Computer simulation is inexpensive, but it may not be reliable. For evaluation of theindoor environment provided by displacement ventilation, the computational fluiddynamics (CFD) technique seems most appropriate, if it is validated by experimentaldata. This combined approach was used to generate the results presented in this

book.

Many experimental data are available in the literature, but very few of them canbe used for the validation. Experimental data for CFD validation must containdetailed information of flow and thermal boundary conditions as well as flow andthermal parameters measured in the space. The data must also include an error anal-ysis. Unfortunately, not many of the experimental data include such detailed infor-mation. Popular data for validating room airflow are from Cheesewright et al. (1986)and Nielsen et al. (1978). Cheesewright's data are for natural convection and

Nielsen's data are for forced convection. However, it is still not certain that a CFDprogram validated by their data can be used for normal room airflow with mixedconvection (a combination of natural and forced convection).

This chapter presents detailed experimental data for displacement ventilation,which is mixed convection and represents ventilation reality in many buildings. Theexperimental data are used to validate a CFD program with a suitable turbulencemodel. There are many turbulence models available. The standard k-model

(Launder and Spalding 1974) is probably most widely used in engineering calcula-tions due to its relative simplicity. However, the model sometimes provides poorresults for indoor airflow. Many modifications have been applied to the standard


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model. However, the modified models do not have a general applicability for indoorairflow. Chen (1995 and 1996) calculated the various indoor flows with eight differ-

ent turbulence models. His study concluded that the re-normalization group (RNG)k-model (Yokhot et al. 1992) is the best among the eddy-viscosity models tested.This chapter will also compare this models prediction for displacement ventilationin a room with the experimental data.

3.1 EXPERIMENTAL FACILITY

The Chambers and HVAC Systems

An environmental test facility was used to study indoor air quality, thermalcomfort, energy efficiency, and ventilation systems. The test facility, shown inFigure 3.1, consists of a well-insulated enclosure. Not shown in the figure are twodoors at either end. A movable wall divides the enclosure into a test chamber and a

climate chamber. We use the larger one as the test chamber and the smaller one asthe climate chamber. The lower part of the movable wall is an insulated exterior wall,and the upper part is a double-glazed window extending almost the whole width.Table 3.1 lists the dimensions and thermal resistance of the chambers.

Each chamber has a separate HVAC system. The two systems are nearly iden-tical. Table 3.1 also lists the capacities of the HVAC systems. Figure 3.2 illustratesthe HVAC system configuration and control interface. Three louvers control outdoorair rate between 0% and 100%. The supply fan and return fan have a variable speeddrive. The interface allows an interactive control of the systems. An operator canchange any parameter, such as airflow rate, supply and return temperature, andhumidity. The air parameters in different sections of the HVAC systems are shownin the monitor and/or are written into a file in the time interval specified by the oper-

Figure 3.1 Sketch of the test facility.


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Experimental Study and Validation of CFD Program

37

ator. Nine-probe hot-wire anemometers, arranged in a matrix form, are used tomeasure the airflow rate. The HVAC and control design allows a variable air supplyrate ranging between 1 ach to 20 ach for the test chamber and 2 ach to 40 ach for theclimate chamber.

Equipment

The major measuring equipment of the test facility includes the following:

A flow visualization system for observing airflow patterns A hot-sphere anemometer system for air velocity, velocity fluctuation, andtemperature measurements

Table 3.1 Dimension, Thermal Resistance, and

HVAC System Capacity of the Test Facility

Length

Width Height Test Chamber Climate Chamber

Dimension Length 17 ft (5.16 m) 10 ft (3.08 m)

Width 12 ft (3.65 m)

Height 8 ft (2.43 m)

Partition heightwindow width

3.9 ft (1.16 m)11.5 ft (3.45 m)

Thermalresistance

Partition wall 30 ft2hF/Btu (5.3 Km2/W)

Partition window 1.5 ft2hF/Btu (0.27 Km2/W)

Other walls 30 ft2hF/Btu (5.3 Km2/W)

Ceiling 30 ft2hF/Btu (5.3 Km2/W)

Floor 30 ft2hF/Btu (5.3 Km2/W)

Door 30 ft2hF/Btu (5.3 Km2/W)

Capacity ofHVAC system

Preheater 8 kW

Supply fan 560 cfm (930 m3/h)

Chiller 21 kW

Reheater 8 kW

Humidifier 11 kgsteam/h none

Return fan 560 cfm (930 m3/h)

Dampers 560 cfm (930 m3/h)


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A tracer-gas system for measuring contaminant concentrations and humidity

A thermocouple system for measuring surface and air temperatures

The test chamber has two long slots and several observation windows on thewalls. Light penetrates through the slots and forms a thin light sheet. By injectingsmoke into the room, the airflow pattern can be observed through those observationwindows normal to the light sheet. The walls of the room were painted black to

enhance visualization. The test facility has two types of smoke source: a theater foggenerator and an air current kit. The theater fog generator produces smoke with atemperature far above the environmental temperature. The smoke is cooled througha long tube before entering the room to approach neutral buoyancy. This generator

produces a large amount of smoke and is good for observing how supply air isdistributed in the room. However, it is not ideal for visualizing airflow pattern in a

particular area. The air current kit remedies this problem. The kit generates a verysmall amount of smoke locally.

Since there is no effective technique at present to measure low air velocitiesthroughout an indoor space, the experiment used hot-sphere anemometers for themeasurements of air velocity, velocity fluctuation, and temperature in the room. AnADD board was used for data acquisition. For velocity, the measurement range of

the hot-sphere anemometers is 10 to 1000 fpm (0.05 to 5 m/s); the repeatability is2 fpm (0.01 m/s) or 2% of the readings. The anemometers cannot reliably measurevelocity when the magnitude is lower than 20 fpm (0.10 m/s). The measuring errors

Figure 3.2 The control interface of the HVAC system.


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for air temperature are 0.8F (0.4 K), including the errors introduced by the dataacquisition systems. Since the probe size is large (about 1/8 in. or 0.003 m in diam-

eter), the probes are not sensitive to high-frequency velocity fluctuation. It is difficultto estimate the errors for velocity fluctuation.

A multi-gas monitor and analyzer system is used to determine indoor air quality.The tracer-gas system can measure many different types of tracer gases. This inves-tigation used SF6and CO2. SF6is better than CO2because the background concen-tration of SF6is almost zero. CO2is inexpensive and was used to check the resultsobtained with SF6.In addition, we also measured the water vapor concentration fordetermining relative humidity in the room. The error for measuring concentration isat an acceptable 10%.

We also used thermocouples to measure air temperature and surface tempera-ture of the room enclosures. The error for measuring temperature by the entiresystem is about 0.8F (0.4 K).

3.2 TEST PROCEDURE

We conducted several measurements with different configurations: a smalloffice, a large office with partitions, and a classroom. Figure 3.3 shows the config-urations used in the experiment.

The internal heat sources and the ventilation rate used in the experiment arelisted in Table 3.2. These cases present some typical situations in the U.S. but theydo not cover a wide range. The main purpose is to obtain experimental data to vali-date a CFD program.

A perforated displacement diffuser, 1.7 ft (0.5 m) wide and 3.6 ft (1.1 m) high,was placed at the middle of the right side wall near the floor in the small office. Theeffective area ratio is 10%. The exhaust, 1.4 ft 1.4 ft (0.43 m 0.43 m), was at thecenter of the ceiling.

The occupants in the test room were simulated by boxes 1.3 ft (0.4 m) long, 1.2ft (0.35 m) wide, and 3.6 ft (1.1 m) high, each heated by three 25 W light bulbs. Themeasured surface temperature was between 82F and 86F (28C and 30C). Two

point sources of SF6were introduced at the top of the two boxes to simulate contam-inants from the occupants, with an initial velocity of 9 fpm (0.045 m/s in the hori-zontal direction). Two PCs were used to generate heat. One generated 108 W and theother 173 W. Six 34 W fluorescent lamps were used during the experiment as over-head lighting. In addition, two tables and two file cabinets were also in the room. Theventilation rate was 4 ach, corresponding to a face velocity of 18 fpm (0.09 m/s) atthe diffuser. The ventilation rate is very high compared to that used in Scandinavia.This is because U.S. buildings have a much higher cooling load based on a surveyin Boston (Chen et al. 1999). A lower ventilation rate will require a lower supply air

temperature. This will lead to draft at ankle level. The supply air temperature corre-sponding to 4 ach was controlled at 62.5F (17.0C). The window surface temper-ature was 81.1F to 82.6F (27.3C to 28.1C), and the surface temperature on the


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(a)

(b)

(c)

Figure 3.3 Space layout used in the experiment: (a) small office, (b) large

office with partitions, and (c) classroom.


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41

movable wall was 75.6F to 79.9F (24.2C to 26.6C). The surface temperatures onthe other walls were 73.9F to 78.8F (23.3C to 26.0C).

Five movable poles were placed in the test room, and each supported six hot-sphere anemometers and six air-sampling tubes. Additionally, two thermocoupleswere also attached on each pole to measure air temperature near the floor and ceiling.A total of 40 thermocouples were used to measure the surface temperatures of thefloor, ceiling, window, and walls.

Measurements were conducted under steady-state conditions by stabilizing theroom thermal and fluid conditions for more than 12 hours before recording the data.Air velocity, air temperature, and SF6concentration were measured in nine different

positions with a total of 54 measuring points for air velocity, 72 for air temperature,

and 54 for tracer gas.The measured air velocities can be used for determining turbulence intensity,

and the measured concentrations can be used to find the local mean age of air. Since

Table 3.2 Specifications for the Experimental Measurements

I-P Units

Case Type

Persons

(number)

Equipment

(Btu/hft2)

Lighting

(Btu/hft2)

Internal

Load

(Btu/hft2)

Ventilation

Rate (ach) Climate

1 Small 2 4.7 3.4 10.7 4 Summer

2 office 2 4.7 3.4 10.7 4 Winter

3 Large 2 4.7 3.4 10.7 8 Summer

4 office 2 4.7 3.4 10.7 4 Winter

5 classroom

6 2.9 3.4 13.9 8 Summer

6 6 0 3.4 11.0 4 Winter

SI Units

Case Type

Persons

(number)

Equipment

(W/m2)

Lighting

(W/m2)

Internal

Load

(W/m2)

Ventilation

Rate (ach) Climate

1 Small 2 14.9 10.8 33.7 4 Summer

2 office 2 14.9 10.8 33.7 4 Winter

3 Large 2 14.9 10.8 33.7 8 Summer

4 office 2 14.9 10.8 33.7 4 Winter

5 classroom

6 9.2 10.8 43.9 8 Summer

6 6 0 10.8 34.7 4 Winter


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42

the omnidirectional anemometers have a large uncertainty in measuring low air

velocity, we feel the fluctuating velocity, |u|, could provide more accurate informa-

tion than turbulence intensity.

This investigation used the step-up and decay method to determine the mean age

of air, :

(3.1)

(3.2)

where c(t), c(0), and c() are the tracer-gas concentration measured at time = t, 0,

and infinity, respectively.

3.3 EXPERIMENTAL RESULTS

Table 3.3 lists the measured air temperatures and concentrations at the supply

inlet and exhaust and the surface velocity at the supply diffuser.

3.4 COMPUTATIONAL FLUID DYNAMICS MODEL

With the detailed experimental data, we can validate a CFD program. A validated

CFD program can then be used to establish a database to evaluate the performance of

displacement ventilation and to develop design guidelines for displacement ventila-

tion. The CFD program used in the present investigation is a commercial program(CHAM 1996) that uses a renormalization group k-model (Yokhot et al. 1992).

Table 3.3 The Measured Parameters

Case Type

Ts TeTracer

Gas

cs ce us

(F) (C) (F) (C) (ppm) (ppm) (fpm) (m/s)

1 Small 62.6 17.0 80.1 26.7 SF6 0 0.42 18 0.09

2 office 63.0 17.2 75.4 24.1 CO2 562 725 18 0.09

3 Large 62.8 17.1 73.6 23.1 CO2 520 602 36 0.18

4 office 62.2 16.8 75.9 24.4 CO2 493 657 18 0.09

5 class-room

67.6 19.8 76.3 24.6 CO2 477 723 37 0.19

6 66.2 19.0 78.4 25.8 CO2 432 924 20 0.10

1c t( ) c 0( )c ( ) c 0( )----------------------------- dt

0

= step-up( )

c t( ) c ( )[ ]dt0

c 0( ) c ( )

--------------------------------------------= decay( )

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