Life Cycle Cost of Displacement Ventilation in an Office Building with a Hot and Humid Climate

LIFE CYCLE COST OF DISPLACEMENT VENTILATION IN AN OFFICE BUILDING WITH A HOT AND HUMID CLIMATE

LANE W. BURT

A THESIS PRESENTED TO THE GRADUATE SCHOOL OF THE UNIVERSITY OF FLORIDA IN PARTIAL FULFILLMENT OF THE REQUIREMENTS FOR THE DEGREE OF MASTER OF SCIENCE

UNIVERSITY OF FLORIDA

2007

To my Dad, whose advice was invaluable

ACKNOWLEDGMENTS

Thanks to my parents for the support they have given me while undertaking this project.

Also, I appreciate the effort of the members of my committee, Dr, Kibert, Dr. Sherif, and especially my advisor, Dr. Ingley. He provided guidance and freedom during the completion of this thesis and reminded me that learning is supposed to be fun and not something you do because they tell you to.

TABLE OF CONTENTS

page

ACKNOWLEDGMENTS ...... 4

LIST OF TABLES...... 7

LIST OF FIGURES ...... 9

ABSTRACT...... 10

CHAPTER

1 INTRODUCTION AND BACKGROUND ...... 11

Motivation...... 11 Displacement Ventilation ...... 12 Application ...... 14

2 LITERATURE REVIEW...... 16

Displacement Ventilation Design Guides...... 16 ASHRAE Design Guidelines ...... 16 REHVA Design Guide ...... 19 Other Literature ...... 20 Humidity Concerns...... 22 Energy Analysis...... 23 Cost Analysis...... 24

3 BUILDING FOR STUDY...... 25

Purpose ...... 25 The Building ...... 25 Specifications ...... 25 Zoning...... 26

4 DISPLACEMENT VENTILATION DESIGN ...... 29

System Design ...... 29 Design Specifications ...... 29 Displacement Ventilation ...... 29 Design Process ...... 29 Application...... 33 Mixing Ventilation ...... 34

5 ENERGY ANALYSIS...... 38

Building Load Profile ...... 38 Bin Weather Data ...... 39 All Zone Level...... 39 System Level ...... 40 Plant Level...... 40 Analysis ...... 41

6 COST ANALYSIS...... 48

Life Cycle Cost Analysis...... 48 Electricity Cost ...... 48 First Costs...... 48

7 CONCLUSIONS ...... 54

Design...... 54 Energy Analysis...... 54 Life Cycle Cost Analysis...... 55 Recommendations for Future Research...... 55

A BUILDING FLOORPLAN...... 57

B DISPLACEMENT VENTILATOR SPREADSHEETS...... 58

Design...... 58 Zone ...... 63 System...... 64 Plant ...... 67 Heating...... 71

C MIXING VENTILATOR SPREADSHEETS...... 74

Design...... 74 Zone ...... 75 System...... 76 Plant ...... 78 Heating...... 81

D WEATHER DATA...... 83

LIST OF REFERENCES...... 87

BIOGRAPHICAL SKETCH ...... 89

LIST OF TABLES

Table page

3-1 Room Zoning ...... 28

4-1 Zone Loads...... 35

4-2 DV Design ...... 36

5-1 Bin Weather Data...... 43

5-2 DV Occupied Zone Level ...... 44

5-3 MV Occupied All Zone Level ...... 44

5-4 DV Occupied System Level...... 45

5-5 DV Occupied Plant Level ...... 46

5-6 Energy Consumption ...... 46

5-7 Checksums ...... 46

5-8 DV and MV Cooling Energy Consumption Comparison...... 47

5-9 DV and MV Energy Consumption Comparison...... 47

6-1 Monthly Electric Demand...... 51

6-2 Electric Rates and Costs...... 51

6-3 First Cost Unit Inputs...... 51

6-4 System First Costs ...... 52

6-5 DV Diffuser Data...... 52

6-6 Life Cycle Cost ...... 53

B-1 Input Parameters for DV Design...... 58

B-2 Output of DV Design...... 58

B-3 DV Design Sizing ...... 59

B-4 DV Occupied Zone Inputs ...... 63

B-5 DV Occupied All Zones...... 63

B-6 DV Unoccupied Zone Inputs ...... 63

B-7 DV Unoccupied All Zones...... 64

B-8 DV Occupied Cooling System Inputs...... 64

B-9 DV Occupied Cooling System Level...... 64

B-10 DV Unoccupied Cooling System Level...... 66

B-11 DV Occupied Cooling Plant Level ...... 68

B-12 DV Unoccupied Cooling Plant Level ...... 69

B-13 DV Occupied Heating Design and Zone Level ...... 71

B-14 DV Heating System Level ...... 72

B-15 DV Heating Plant Level...... 73

C-1 MV Design Level...... 74

C-2 MV Design Output...... 75

C-3 MV Occupied Cooling Zone Level...... 75

C-4 MV Unoccupied Cooling Zone Level...... 75

C-5 MV Occupied System Level...... 76

C-6 MV Unoccupied Cooling System Level...... 77

C-7 MV Occupied Cooling Plant Level...... 79

C-8 MV Unoccupied Cooling Plant Level...... 79

C-9 MV Heating Design and Zone Level...... 81

C-10 MV Heating System Level...... 82

C-11 MV Heating Plant Level ...... 82

D-1 Monthly Design Day Weather Data...... 83

LIST OF FIGURES

Figure page

1-1 Mixing vs. Displacement Ventilation ...... 15

3-1 First Floor Plan ...... 27

5-1 Building Load Profile ...... 43

A-1 Building Floorplan...... 57

B-1 Part Load Operation Data for 40 ton Trane Chiller ...... 67

B-2 Effect of Ambient Temperature on 40 ton Trane Chiller ...... 68

C-1 Part Load Operation Data for 50 ton Trane Chiller ...... 78

C-2 Effect of Ambient Temperature on 50 ton Trane Chiller ...... 78

Abstract of Thesis Presented to the Graduate School of the University of Florida in Partial Fulfillment of the Requirements for the Degree of Master of Science

LIFE CYCLE COST OF DISPLACEMENT VENTILATION IN AN OFFICE BUILDING WITH A HOT AND HUMID CLIMATE

Lane W. Burt

December 2007

Chair: H. A. Ingley Major: Mechanical Engineering

Building energy use is the subject of increased scrutiny as the costs and environmental impacts of energy use continue to rise. The building mechanical systems represent a significant portion of building energy use and should be analyzed to determine best practices. Indoor air quality is also a concern related to the mechanical system; however, energy efficiency and indoor air quality are often competing forces in the building design process.

Displacement ventilation is a strategy that has been proven to provide superior air quality in European buildings and may possibly save energy. The strategy has not been implemented widely in the US and not at all in a hot and humid climate. Further, energy performance and cost estimates are not available for most applications. This study simulates the energy performance of a displacement ventilation system in an office building in Gainesville, FL using a modified bin method and then completes a life cycle cost analysis to compare the displacement ventilation system to a traditional mixing ventilation system. The results show that a displacement ventilation system in Florida would use slightly more energy than a traditional system with a substantial first cost premium. The overall life cycle cost of the displacement system would be higher than the mixing system.

CHAPTER 1 INTRODUCTION AND BACKGROUND

Motivation

As energy prices continue to rise, so does the scrutiny and critical examination of energy use. Overall, more than one-third of US energy is consumed by buildings (Chen and Glicksman

2003), as well as two thirds of all electrical energy (DOE 2000). In environmental terms, 50% of

all CO2 emissions from industrialized countries are associated with building operations (Loveday et al 2004). These figures along with the 20 to 100 year service lives of buildings show why so much attention should be paid to building energy consumption.

Energy efficiency is not the only parameter for a well designed building. The building first

must be comfortable, as many studies have shown that productivity and satisfaction are directly

related to comfort. Most recently, Budaiwi found in 2005 that undesirable thermal conditions

have an adverse effect on productivity. Thermal comfort is now better understood and thus new

buildings can be designed and constructed to be extremely appealing.

Beyond comfort, the building must provide a safe and clean indoor environment. Modern

people spend a majority of their time indoors. Jenkins et al (1992) found that the mean

percentage of the day spent inside for the average person was 87%. Seppanen and Fisk (2004)

surveyed the existing literature and determined that ventilation has a significant effect on task

performance and productivity. The study of indoor air quality (IAQ) became important when

building envelope design began to restrict ventilation air flow to the point that pollutants began

to accumulate inside, causing the first diagnoses of ‘sick building syndrome.’

By considering efficiency, comfort, IAQ, and first cost, building design has improved

tremendously in the time since the energy crisis of the late seventies. New materials, techniques,

and computer modeling programs have made buildings more efficient, more comfortable, and

cleaner; however, improvements can still be made. New ideas are being tested and old ideas are being re-examined using modern knowledge. Strategies and expertise sharing around the world is leading to new ways to implement proven technologies and produce better performing buildings. For example, a proven technology in Europe, displacement ventilation, may have applications in a geographical area such as the US and more specifically Florida and may provide new insights into improving building performance.

Displacement Ventilation

Displacement ventilation (DV) is the process of supplying air into a room at a temperature slightly lower than that of the desired ambient temperature and at a slow speed. The air will flow along the floor and will rise by natural convection when it heats up or when it comes into contact with a heat source, such as a piece of equipment or a person. After contact with a heat source, the air will be warmed even faster, and will rise up the surface of the heat source towards the ceiling. There will be a vertical temperature gradient created in the room and temperature stratification will occur. Two stratified zones are normally identified – the breathing zone, found from the floor to the head level of the occupant and the contaminant zone, found from the ceiling down to the top of the breathing zone. For the occupant, DV provides comfortable, high quality air in the breathing zone. The warm contaminated air will then be exhausted at the ceiling. This system is extremely different from traditional mixing distribution, or dilution ventilation. In mixing ventilation (MV), air is supplied to a room at high velocity from the ceiling and all the air in the room is mixed together to provide a uniform temperature. The downside to this approach is that contaminants are mixed with the air where they are continuously present in the breathing zone; thus, the indoor air quality is not as high as in displacement ventilation. Figure 1-1 illustrates the differences between MV and DV.

The bulk of the research on displacement ventilation has been conducted in Scandinavian

countries, where it has been used commonly since its introduction in 1978. In Nordic countries in 1989, DV had a 50% market share for industrial applications and 25% for offices (Svensson,

1989). Other European countries have similarly adopted DV to ventilate and cool their buildings, resulting in the publication of a multi-national guidebook in 2001.

There are fundamental differences in the motivation to apply DV in the US and Europe. In

Europe, DV is seen as a way to improve indoor air quality while having little impact on the

energy use. Americans view DV as a way to maintain acceptable or improved IAQ while

reducing energy consumption. These different goals are also affected by physical differences in

the buildings. US buildings tend to have higher cooling loads than those found in European

buildings, due to the warmer climate and the abundance of internal loads generated by lighting

and equipment. The cooling of perimeter areas of US buildings is driven by envelope loads,

while internal loads dominate the core areas. Perimeter spaces need heating or cooling

depending on outdoor conditions, while core spaces may need constant cooling as a result of

internal loads. Northern European buildings deal with this issue by using radiant heating in the

winter and displacement ventilation to cool in the summer. DV is not suitable for heating

because a high supply air temperature causes a heating short circuit as warm air rises quickly

after being supplied to the space and is exhausted directly. The radiator heats the space and

surfaces and insures that the supply temperature of the ventilation air is always cooler than the

ambient temperature, and thus buoyancy driven flows occur and the ventilation is still effective.

US buildings traditionally utilize the same system for heating and cooling, which may cause a

heating short circuit in the summer. If warm air were introduced into the space using the

displacement system, its natural buoyancy would cause it to rise and be exhausted from the space before heating could occur, making using a DV system for heating and cooling unwise.

Application

The use of DV in some applications is well documented, mainly in larger rooms like amphitheaters, industrial spaces, and classrooms, while other applications are not as well understood, such as offices. More specialized building types such as cleanrooms or laboratories have never been considered using DV. In the US, application is starting slowly with classrooms and computer rooms while interest in office building implementation is growing.

The variation of the US climate also complicates the DV application process as implementation may be beneficial in some areas and not in others. Lines of feasible application have never been mapped out although such research has been called for. Applications in

California, the Pacific Northwest and Canada have been aided by mild climate similar to

Northern Europe. In dissimilar areas, such as hot and humid Florida, DV has received little or no attention. Application in these areas is discouraged by a lack of information, making benefits and drawbacks completely unknown. A further look into these issues is needed.

Overall, there is a lack of field data on DV, as most studies are simulation only. This is clearly due to a lack of buildings using the technology. Similarly, most building simulations are conducted on a single unit basis and do not consider the interaction between rooms that will take place in most indoor environments. Also, cost data on these systems are largely unavailable.

There exists only an overall assumption that DV systems are more expensive than traditional mixing systems, but this has yet to be proven. For these reasons, this project seeks to analyze an office building using a larger than single-unit framework, in Florida, and complete a Life Cycle

Cost (LCC) analysis to put the results into financial terms.

Figure 1-1 Mixing vs. Displacement Ventilation

CHAPTER 2 LITERATURE REVIEW

Displacement Ventilation Design Guides

Two major publications have been produced to explain how to design displacement

ventilation systems. The American Society of Heating, Refrigerating, and Air-Conditioning

Engineers (ASHRAE) produced System Performance Evaluation and Design Guidelines for

Displacement Ventilation for applications in the US in 2003 written by Chen and Glicksman.

The European counterpart is the Federation of European Heating and Air-Conditioning

Associates’ (REHVA) publication Displacement Ventilation in Non-Industrial Premises, which

was made available in 2001. These design guides seek to simplify the design process which

previously used energy balance methods on room by room levels that are far too complicated for

manual calculations, as noted by Livchak and Nall (2001)

ASHRAE Design Guidelines

The ASHRAE publication contains a literature review, a CFD experimental study with

validation, explanations of temperature and ventilation models, performance evaluation criteria,

energy and cost analysis, and finally step by step design guidelines. The guide seeks to illustrate

the differences between the relatively new application of displacement ventilation to US

buildings and the mature applications found in Europe.

The literature review in the ASHRAE guide is extensive and this survey will focus on important developments in the document and later will note new research. The guide explains the development of DV in terms of temperature distribution, flow distribution, contaminant distribution, and thermal comfort.

Temperature Distribution: The vertical temperature gradient is important in evaluating

DV, as a gradient that is too large will cause drafts and not provide thermal comfort to the

occupants. The vertical temperature gradient is a function of the ventilation rate, while heat

source type, internal convection and radiation, space height, and diffuser type all play a role.

The current methods do not take into account all of these factors and thus may provide false

values. A calculation procedure showing the influence of all of these factors needs to be created.

Flow Distribution: The stratification height, which is the height at which contaminants are no longer carried upwards by the DV system, must be located above the breathing zone. This height is a function of the supply air velocity and the location, geometry, and strength of the thermal plumes generated by the heat source. This height is affected by the vertical temperature gradient. Diffusers can cause draft if air supply rates are too high; however, manufacturer’s data should be sufficient to avoid this problem.

Contaminant Distribution: The concentration of contaminants in the breathing zone is

dependent on the type of contaminant, the location, and the proximity of heat sources. If the

contaminants are associated with heat sources there will be a low concentration of the

contaminant in the breathing zone. The air will make contact with a person, be warmed and rise

by natural convection into the breathing zone. This air will be of higher quality than the ambient

air at that level and the mean age of air (the time since the air was introduced into the space) will

be lower. Cool surfaces can increase the concentration of contaminants in the breathing zone.

There is not a sufficient method for the accurate prediction of contaminant concentration, but

consensus research shows improved IAQ when using DV over ceiling based mixing systems

Thermal Comfort: Drafts and temperature gradient represent the largest risks to thermal

comfort when using DV. The temperature gradient can cause discomfort and can be reduced

with increased supply air flow, but this increase causes a rise in draft risk as well as increased

energy consumption. The previous research points to a cooling load limit of 13 BTU/hr*ft2 (40

W/m2) but this limit may be much higher and thus the application of DV may be extended.

ASHRAE concludes that the cooling load limit should be 38 BTU/hr*ft2 (120 W/m2)

The ASHRAE guide book also attempts to point out the problems associated with previous work. The vertical temperature gradient is not constant, as assumed in the calculation procedure, and thus may cause the selection of higher ventilation rates than is actually necessary. Because of this and other assumptions made, as well as a survey of designers, the authors conclude that previous design guidelines cannot be used with confidence.

There is also a validated model of DV using Computational Fluid Dynamics (CFD) in the

ASHRAE book that concludes that a properly designed DV system will maintain a higher indoor

air quality than a mixing ventilation system while providing acceptable comfort conditions. The

CFD model was validated using a test chamber set up to mimic a small office, cubicle office,

quarter of a classroom, and a workshop. Air flow, air temperature, tracer gas concentration, air

velocity, and the mean age of air were considered. Tracer gas measurements did not agree with

the model and thus contaminant concentration modeling is not trustworthy. The CFD model was

then used to create temperature difference prediction models as well as ventilation effectiveness

models and performance evaluation criteria.

The step by step design process outlined in the ASHRAE book is as follows,

1: Judge the applicability of DV 2: Calculate the summer design cooling load 3: Determine the required flow rate of the supply air for summer cooling 4: Find the required flow rate of fresh air for acceptable indoor air quality 5: Determine the supply flow rate 6: Calculate the supply air temperature 7: Determine the ratio of the fresh air to the supply air 8: Select the air diffuser size and number 9: Check the winter heating situation 10: Estimate the first costs and annual energy consumption

This design process will be followed in this study and further explanation of each step can be

found in the next chapters.

REHVA Design Guide

The REHVA design guide was the immediate predecessor of the ASHRAE design guide.

The guide is mostly in agreement with the ASHRAE guide, and the differences will be

highlighted here.

The REHVA design guide is more extensive than the ASHRAE version, consistently providing more information on the development process. The processes determining air flow

characteristics, temperature models, plume development, contamination distribution, ventilation

effectiveness, and diffuser design are all detailed. Further, control strategies are also discussed.

Rather than supplying step by step design guidelines like the ASHRAE guide, REHVA presents five case studies of different building types and walks the user through the design process undertaken in that building type to provide an example. The office case study is, like ASHRAE, a single cell unit. Along the way, the guide gives various warnings at points where the designer

could make a mistake and cause drafts or temperature gradient problems in the room, as well as

common mistakes in diffuser selection.

The design procedure recommended by REHVA is a completely different approach from

ASHRAE. The differences begin at step one, as the goal of the system must first be specified in

air quality or temperature. Temperature design means that the system will primarily exist to

remove large heat surpluses with large quantities of air. Air quality design sets out to meet a

given level of IAQ through ventilation. The different design goals will result in extremely

different systems. The design guidelines for air quality are as follows,

1: Select the stratification height 2: Determine the convective flow rates through the stratification height 3: Choose the supply air flow rate 4: Calculate the exhaust contaminant concentration in the occupied zone 5: Evaluate the concentration in the occupied zone 6: Check that the air flow rate is sufficient according to codes and standards 7: Choose the air volume flow with regards to temperatures, air quality, and regulations 8: Recalculate the vertical temperature distribution in the room and estimate the pollutant stratification height 9: Select diffusers and ensure that the adjacent zones are acceptable

The design guidelines for temperature control, which are comparable to the ASHRAE guidelines, are as follows,

1: Select thermal comfort criteria 2: Calculate the heat surplus to be removed by the ventilating air 3: Calculate the maximum temperature increase from supply to exhaust air 4: Calculate the supply air temperature 5: Calculate the supply air volume flow rate 6: Re-evaluate the air temperature increase at the floor level 7: Check that the air flow rate is sufficient according to codes and standards 8: Choose the air volume flow with regards to temperatures, air quality, and regulations 9: Recalculate the vertical temperature distribution in the room and estimate the pollutant stratification height 10: Select diffusers and ensure that the adjacent zones are acceptable

The REHVA design guide functions in terms of temperature gradients and contaminant distribution, which ASHRAE does not deal with directly in their design procedure. REHVA also mentions the interaction between adjacent spaces, which is not covered at all by ASHRAE. The space cooling loads considered acceptable by REHVA are much lower than those acceptable to

ASHRAE (13 compared to 38). Because the ASHRAE guide was written to deal with the higher loads of the US, it will be followed in this study. Also, the building does not meet the cooling load limits established by REHVA.

Other Literature

The National Institute of Standards and Technology (NIST) conducted an initial evaluation of the use of DV in US commercial buildings in 2005 (Emmerich and McDowell). This

document surveyed existing literature, including the two guidebooks, and evaluated the positives and negatives associated with DV when specifically considering US application. The document strives to note the weaknesses of DV that may have been glossed over by previous work, which are,

• Care must be taken to balance the needs of IAQ and thermal comfort (drafts, temperature gradients)

• Stratification may be affected by localized room conditions, resulting in the lowering of the contaminants into the breathing zone

• Contaminants not associated with heat sources may not be removed effectively as most research has focused on contaminants that are associated with heat sources

• Occupant movement, cross ventilation or other disturbances may reduce DV to MV like conditions

• Humidity conditions are relatively unknown

These concerns are being addressed by current research and NIST notes the specific projects dealing with each point. Of particular importance is the demonstration that thermally neutral contaminants in the occupied space show the same concentrations in DV as in MV systems, meaning worst case scenario for a DV system is mixing. Also, occupants can similarly cause DV systems to act as mixing through motion. Nielsen et al (2003) determined that increasing supply air reduces the predicted percentage dissatisfied (PPD meaning the number of people who will complain in a given condition) due to drafts, while decreasing the supply air flow reduces PPD due to vertical temperature differences. An optimal airflow rate for both parameters exists, but provides a very narrow range of acceptable flow rates. Also, the actual thermal comfort performance of DV systems has not been proven. Melikov et al (2005) surveyed occupants in a DV building and noted that nearly one half were made uncomfortable daily due to temperature and one quarter due to drafts. This was not compared to an MV survey.

Clearly, more research is needed. Also, NIST correctly notes that humidity is an important parameter to take into account when considering the use of DV in a hot and humid climate such as Florida.

Humidity Concerns

Research into the interaction of room humidity levels in displacement ventilation systems is relatively new. The guidebooks do not deal explicitly with this topic. Since implementation has taken place mostly in mild or cooler climates such as Northern Europe, humidity was not a large concern. In humid areas, outside air must be dehumidified before it can be brought into the space, eliminating much of the prospective energy savings found in DV, which often uses large quantities of outside air. Northern European systems often use 100% outdoor air. Livchak and

Nall (2001) note that this difference means that implementation and control of DV in humid areas must be different.

NIST (2005) details work into humidity levels in DV and draws the following conclusions,

• Humidity levels stratify with vertical temperature when using DV

• Humidity stratification is significant and may be an extra energy benefit to use DV in humid areas, although some humidity levels may be high enough to be of concern

• Because of humidity, occupants are comfortable at slightly higher temperatures when using DV than MV

Other studies into humidity levels have taken place mainly in hot and humid areas of Asia.

Kosonen (2002) studied a factory in Malaysia and found significant humidity stratification levels. These measured humidity levels were less than those predicted by the model and caused an over sizing of the cooling equipment by 6% and thus improving models may lead to energy savings. Cheong (2004) similarly found that cooling capacity for DV should be 5% lower than

MV after taking into account humidity factors. Raising temperatures 1°C (1.8 °F) also resulted in energy savings, as sedentary occupants feel cooler sensations when in a DV environment.

To control the humidity of the space, Livchak and Nall (2001) recommend recirculating room air after it passes the cooling coil. The conclusion was made that this strategy allowed the system to control supply air moisture without increasing cooling capacity. This recirculation will also have an impact on the energy consumption of the system. This is not the case in northern

Europe, as Hensen and Hamelinck (1995) stated that in the climate of the Netherlands recirculating air has no impact on energy and is therefore not advisable. Thus this factor has been ignored and new design strategies for humid areas must be developed. Overall, humidity control issues have not been explored sufficiently for special measures to be taken in this study.

Energy Analysis

There has not been much work in energy analysis of DV systems. The ASHRAE design guide notes that,

• Previous research has been by numerical simulation rather than actual measurements

• Energy findings have been too varied to conclude that DV does or does not save energy

• Application to core spaces works, but perimeter spaces may not be suitable for the system because of high cooling load

• DV has often been combined with other systems such as cooled ceiling panels, which provides a higher first cost than mixing systems

In the energy study conducted in the ASHRAE design book, five climate regions were considered with New Orleans, LA representing hot and humid areas. The study focused only on single units with various configurations of walls and windows. The conclusion was made that the DV systems used more fan energy and less chiller and boiler energy than MV. Total energy for DV was slightly less than MV.

Modeling DV systems is not at all a developed practice. Current energy modeling packages do not have the ability to consider vertical temperature differences and thus all computer models using these packages make assumptions to work around these weaknesses. A model for use in the simulation program EnergyPlus was created by Carrilho da Graça in 2003 but its implementation is limited to certain room geometries and accuracy is not quantified.

Livchak and Nall (2001) used the bin method to compare the energy performance of DV to

MV in a hot and humid climate, using return air recirculation. They concluded that displacement ventilation systems allow reducing the cooling capacity and the annual energy consumption by the chiller for the high heat load applications. Because of the lack of energy modeling software and the previous use of the bin method in past studies, it will be used in this study.

Cost Analysis

Little effort has been made to estimate costs of DV in comparison to MV. Assumptions are normally made that alternative strategies will cost more money, but not much data exist to support or disprove this idea. NIST (2005) notes that additional work is needed to determine if first cost and operating cost advantages exist along with IAQ advantages when comparing DV to

MV. The ASHRAE design guide attempts to estimate first costs and determines DV saves first cost by downsizing the chiller but increases cost by over-sizing the air handler. Air distribution products such as duct and diffusers were not considered. The conclusion was made that first cost for DV systems is slightly higher than MV systems because of the separate heating system needed.

Life Cycle Cost Analysis (LCCA) will be used to attempt to estimate the costs of a DV system. This process will be based upon NIST Handbook 135, which is the standard manual of

life cycle costing written by Fuller and Peterson in 1995.

CHAPTER 3 BUILDING FOR STUDY

Purpose

The goal of this research was to select an existing building in Florida that represents

normal MV application. The building mechanical system would then be re-designed using DV.

Both systems would be modeled to show their energy consumption and finally, LCC would be

used to compare the costs of a DV system in a hot and humid climate to a traditional MV system.

The Building

The building chosen is located in Gainesville, FL and is three stories with a penthouse and ceiling heights of 12 ft and floor to floor dimensions of 16ft. The building utilizes two 125 ton chillers for the three floors, while a separate unit serves the penthouse. A VAV system with supply and exhaust at the ceiling level is used for air distribution. The first floor is an open office concept while the second has more individual offices. To simplify the modeling process, only the first floor is considered in this project. Furthermore, specialized rooms such as the lobby, mechanical, and electrical rooms were also not included. Data available from the building include,

• A full set of as built drawings • A usable CAD file • The original load calculations from the designer

The first floor building plan is shown in figure 3-1. A full plot can be found in Appendix A.

Specifications

The building is constructed of concrete masonry units. The U factor for the walls was

assumed to be 0.11 BTU/hr*ft2*°F (0.62 W/m2*C) based on the original load calculation from

the design engineer. The U factor of the floor was 0.21 BTU/hr*ft2*°F (1.21 W/m2*C). The

glass possesses very high performance attributes with a U factor and shading coefficient of 0.35.

Zoning

The building rooms were divided into zones, with each zone served by a VAV box. The zones can be seen in table 3-1. The naming system used in the zone varies, as the individually

numbered zones were created by the author, while the ‘1-##’ zones are identical to the actual

zoning of the building and were therefore allowed to keep the same name. The non numbered

zones represent large rooms that contain multiple zones. The same zoning was used for the DV

and MV systems.

Figure 3-1 First Floor Plan

Table 3-1 Room Zoning Zone Zone Name Room Load Name Rm # Name Room Load Name Rm # 3 Corridor 2 144 13 Workroom 103 3 Board Room 117 13 R Support 102A 2 Toilet 110 17 Receiving 148 2 Coffee 116 17 Corridor 4 ext 149 16 Corridor 4 int 154,156 21 Marketing West 137 16 Coffee Bar 147 22 Marketing East 137 12 Corridor 3 163 23 Research Lib West 159 12 Comp Dept Office 160 24 Research Lib Mid 159 12 Comp Dept Cubicles 161 25 Research Lib East 159 5 Accounts Payable 126 19 Lounge 146 5 Consultant 124 20 Meeting Room 145 1 Corridor 1s 133 18 Unassigned Office 157A 1 Exec Asst 107, 112, 108, 111 18 Unassigned Office 157 4 Exec Asst 121, 122 18 Storage 156 4 Storage 123 18 Research Office 158 9 Human Resources 138 1-14 Marketing 134 9 Human Resources 139 1-14 Marketing 135 6 Interview 143 1-14 Dir Marketing 136 6 Recruiter 142 1-13 Dir Edit Services 131 7 Recruiter 141 1-13 Dir Hmn Resources 129 7 Human Resources 140 1-11 Dir Inside Sales 127 10 Unassigned Office 132 1-11 Dir Outside Sales 125 10 Storage 130 1-08 VP Sales 120 8 Corridor 1n 133 1-06 CFO 119 8 Unassigned Office 128 1-05 COO 118 14 Men's 164, 165 1-04 President 115 14 Custodial 166 1-03 Mtg Rm & Dining 113 14 Women's 167, 168 1-01 Chairman 106

CHAPTER 4 DISPLACEMENT VENTILATION DESIGN

System Design

Design Specifications

Both systems were designed with the following criteria,

• 75°F (23.9 C) cooling setpoint, 68°F (20 C) heating setpoint • 78°F (25.5 C) cooling drift-point, 55°F (12.8 C) heating drift-point • 57°F (13.9 C) cooling coil leaving dry bulb temperature (for dehumidification) • ASHRAE 62.1 compliant outdoor air ventilation rates

Displacement Ventilation

Design Process

The DV system was designed using the ASHRAE design guidelines. All equations are

from ASHRAE. The following steps were followed,

Judge the applicability of DV: In this building, the ceilings are over 8 ft tall and are physically

suitable for DV.

Calculate the summer design cooling load: A load calculation program was used to determine the loads. The same loads were used in both the DV and the MV system. The heat gains within the occupied zone are itemized as follows:

QOE = heat gains from people, equipment, and task lighting

QL= heat gains from overhead lighting

QEX = heat gains from envelope loads (ceilings, infiltration, solar heat gains, etc.)

The max load per square foot of floor area was calculated. This guide suggests that DV

applications can handle up to 38 BTU/hr*ft2 (120 W/m2) of cooling. ASHRAE states that any

cooling-load density greater than this value is not suitable for DV and additional technologies

such as chilled beam systems may be used. Other guides have lower cooling load limits.

Calculate the air changes per hour: Based on heat gains using equation 4-1,

295.0( QOE + QL + QEX )185.0132.0 n = (4-1) Δ ρ p HACT

With ΔT = delta temperature between foot and head level (taken to be 3.6F or 2C)

ρ = density of air (lb/ft3)

Cp = specific heat of air (BTU/lb*F)

H = height of ceiling (ft)

A = floor area (ft2)

n = ventilation rate in air changes per hour (acph)

The coefficients used in front of the heat gains represent the fractions of the cooling load

entering the space between the head and feet of a sedentary occupant, meaning only the occupied

zone is being condition, which is approximately six feet (1.83 m) up from the ground and a foot

(0.30 m) from each vertical wall.

The ventilation rate, Vc, required for summer cooling was then calculated in CFM from

equation 4-2,

nHA V& = (4-2) C 60

Calculate the flow rate of fresh air: This value is calculated by using the ASHRAE 62.1 2004 required flow rate and the ventilation effectiveness, η. The actual effectiveness can be calculated as follows in equation

4-3,

QOE + QL + QEX )5.04.0( η −= e− 28.0 n )1(4.3 (4-3) ( ++ QEXQLQOE )

The ventilation effectiveness is a measure of how much fresh air is needed to attain acceptable indoor air quality. ASHRAE Std. 62-2001 states that typical ventilation effectiveness for a DV system is 1.2 while mixing is assumed to be 1.0.

The outdoor air ventilation rate for the DV system can then be calculated from equation

4-4,

V 1.62 VOA = (4-4) η

With the variable V62.1 representing the ventilation rate needed for the occupied space based on ASHRAE Std. 62-2004 and VOA is the actual amount of outdoor air needed by the DV system.

Choose flow rate: Based on either outside air rate or cooling load rate, as shown in equation 4-5,

V = VCVOA ),max( (4-5)

The air flow rate is based on the maximum value between the summer cooling flow rate and the flow rate for IAQ. If the fresh air value, VOA is greater then the room will operate on

100% OA.

Calculate the supply air temperature: The dimensionless temperature, θf, can be calculated by equation 4-6,

−1 ⎡ ρCV p )60( 11 ⎤ θ f = ⎢ (* ++ 1) ⎥ (4-6) ⎣ A R αα CF ⎦

With αR = radiative heat transfer coefficient

αCF = convective heat transfer coefficient

Both coefficients have units of BTU/(hr*ft2*°F) and are approximately 1 as stated by

ASHRAE. The nrequired supply temperature based on dimensionless temperature can be calculated from equation 4-7,

θ f (( + + QEXQLQOE )) SADB ( TTSET hf ) −Δ−= (4-7) 60ρ pVC

With TSET = design room temperature (i.e. 75 °F, 23.9 C)

ΔThf = delta temperature from head to foot level (3.6 F, 2 C)

The supply air temperature (SADB) must be calculated for the air handler. This temperature is selected as the maximum supply temperature found in all the zones being considered. It should be noted that these room temperatures will interact unless the rooms are closed off from each other. The consequences of these interactions are not quantified. The ventilation rates for the non maximum rooms are now inaccurate and they have to be recalculated using this supply temperature. ASHRAE notes that these ventilation rate calculations are not very accurate.

The return air temperature (RADB) can be calculated from the energy balance in equation

4-8

Q RADB SADB += ∑ (4-8) 60ρ pVC

Determine the ratio of fresh air to supply air: The percentage of outside air, XOA, can then be calculated for all rooms in equation 4-9,

VOA XOA = (4-9) V

Typical office buildings utilize between 15 to 30% outside air based on the number of people within the occupied space, building codes, etc.

Calculate diffuser face area: Area, in equation 4-10,

= 40/ fpmVA (4-10)

The diffuser face area is based on the supply air flow rate and the maximum velocity allowed by the diffuser. ASHRAE concludes that the maximum flow rate for DV applications is

40 feet per minute (fpm) (0.2 m/s). Higher rates are thought to cause draft with current diffuser design.

Check the winter heating situation: Since DV is not usually suitable for heating because of the short circuit effect already described, some sort of separate heating system will have to be specified.

Estimate first costs and energy consumption: The guide does not give specific methods for completing this step, only noting that the previous work on building types can be used.

Application

To apply this design procedure to the building selected, the loads were first calculated on a per square foot basis to check applicability to ASHRAE standards. The load density of the original design was too great in several rooms along the perimeter of the building, so overhangs were added to reduce the exterior loads. The same loads were used for both the DV and MV models, so overhangs were included in both. The loads are shown in table 4-1.

The loads are all under 38 BTU/hr*ft2 (120 W/m2) and the ceiling height is 12 feet (3.6 m), making these zones suitable for displacement ventilation by ASHRAE standards. The design sizing of the DV system was completed based on these guidelines, and the full spreadsheet can be found in appendix B. An abbreviated design sizing spreadsheet can be found in table 4-2.

From the design process it is determined that the supply temperature (SADB) for the DV system is 68.7°F (20.4 C), the average RADB is 83.4°F (28.5 C) and the system must operate on at least

17% outdoor air.

Mixing Ventilation

The mixing ventilation system was designed as a typical VAV system. All relevant design specifications were kept exactly the same as the DV system. The full design spreadsheet can be found in appendix C. The SADB for this system was 57°F (13.9 C), the average RADB was

76.2°F (24.5 C) and the system operates on at least 19% outdoor air.

Table 4-1 Zone Loads QL QEX QOE Area (lights) (exterior) (internal) ∑Q Q/A Zone (ft2) (BTU/hr) (BTU/hr) (BTU/hr) (BTU/hr) (BTU/ft2*hr) 3 894 4,638 0 6,498 12030 13.5 2 213 1,105 0 3,242 4560 21.4 16 644 3,341 0 6,485 10470 16.3 12 723 3,751 0 3,295 7769 10.7 5 258 1,338 0 2,197 3793 14.7 1 525 2,724 0 1,548 4797 9.1 4 308 1,598 0 2,521 4427 14.4 9 252 1,307 0 2,197 3756 14.9 6 227 1,178 0 2,197 3602 15.9 7 257 1,333 0 2,197 3787 14.7 10 194 1,006 0 1,098 2299 11.9 8 405 1,006 0 2,197 3608 8.9 14 659 3,419 0 0 4078 6.2 13 300 1,556 0 7,060 8917 29.7 17 285 1,401 2,440 900 5026 17.6 DV Marketing West 582 2,860 2,177 5,322 10941 18.8 DV Marketing East 640 3,145 3,019 4,257 11061 17.3 DV Research West 962 4,728 9,939 2,579 18208 18.9 DV Research Mid 546 2,385 7,247 720 10898 20 DV Research East 831 4,311 20,248 2,197 27587 33.2 19 1315 6,463 6,436 10,800 25014 19 20 515 2,531 2,286 6,464 11797 22.9 18 641 2,800 9,922 2,988 16351 25.5 1-14 525 2,259 13,118 2,988 18889 36 1-13 308 1,346 6,106 1,992 9751 31.7 1-11 308 1,346 6,106 1,992 9751 31.7 1-08 227 992 4,499 1,446 7164 31.6 1-06 227 992 4,499 1,446 7164 31.6 1-05 227 992 4,499 1,446 7164 31.6 1-04 255 1,114 5,061 1,446 7876 30.9 1-03 384 1,678 6,449 2,346 10857 28.3 1-01 306 1,474 4,636 1,514 7931 25.9 Total 14,943 72,116 118,687 95,579 301325

Table 4-2 DV Design Cooling Effective- Outdoor Outdoor Temp Supply Flowrate ness Air Air DV Flowrate Gradient Temp Zone n Vc η V62.1 VOA V θf SADB Name (ACPH) (cfm) (cfm) (cfm) (cfm) (F) 3 3.64 651 1.63 96 59 651 0.388 65.2 2 6.66 284 2.44 18 7 284 0.258 67.7 16 4.71 606 1.98 44 22 606 0.329 66.4 12 2.30 332 1.05 58 56 332 0.502 62.4 5 3.53 182 1.59 25 16 182 0.396 65.1 1 1.71 180 0.75 42 55 180 0.575 59.9 4 3.50 215 1.58 17 11 215 0.398 65.0 9 3.59 181 1.61 25 16 181 0.392 65.2 6 3.89 177 1.72 24 14 177 0.373 65.6 7 3.54 182 1.59 25 16 182 0.395 65.1 10 2.64 102 1.21 12 10 102 0.467 63.3 8 2.11 171 1.19 29 25 171 0.523 63.4 14 0.88 116 0.30 16 53 116 0.724 51.6 13 9.32 559 2.79 8 3 559 0.199 68.7 17 3.54 202 1.12 5 5 202 0.395 63.5 21 4.66 542 1.75 60 34 542 0.332 66.0 22 4.09 524 1.51 58 39 524 0.361 65.2 23 4.11 790 1.24 130 105 790 0.360 64.3 24 4.18 457 1.16 76 65 457 0.356 64.1 25 7.51 1248 1.55 110 71 1248 0.235 66.8 19 3.73 982 1.38 104 75 982 0.382 64.6 20 4.77 491 1.74 88 50 491 0.327 66.0 18 5.70 731 1.50 34 23 731 0.289 65.9 1-14 8.03 843 1.65 47 28 843 0.223 67.2 1-13 7.00 431 1.62 28 18 431 0.248 66.7 1-11 7.00 431 1.62 28 18 431 0.248 66.7 1-08 6.63 301 1.55 19 12 301 0.259 66.5 1-06 6.63 301 1.55 19 12 301 0.259 66.5 1-05 6.63 301 1.55 19 12 301 0.259 66.5 1-04 6.51 332 1.52 20 13 332 0.262 66.4 1-03 6.37 490 1.60 43 27 490 0.266 66.5 1-02 5.85 358 1.53 23 15 358 0.283 66.1 Totals & Averages 985 13693

Table 4-2 Continued V Return OA recalc Temp Fraction (cfm) RADB XOA Diffuser 2) Zone Name (F) Area (ft 3 832 81.1 7% 20.8 2 304 81.9 2% 7.6 16 713 81.5 3% 17.8 12 512 80.4 11% 12.8 5 235 81.0 7% 5.9 1 321 79.9 17% 8.0 4 279 81.0 4% 7.0 9 232 81.1 7% 5.8 6 221 81.2 6% 5.5 7 235 81.0 7% 5.9 10 148 80.6 7% 3.7 8 261 78.7 9% 6.5 14 310 78.9 17% 7.7 13 559 82.3 1% 14.0 17 273 83.5 2% 6.8 21 650 82.3 5% 16.2 22 658 82.4 6% 16.5 23 1016 83.9 10% 25.4 24 588 84.5 11% 14.7 25 1382 86.4 5% 34.6 19 1277 82.4 6% 31.9 20 586 82.5 9% 14.7 18 855 84.8 3% 21.4 1-14 917 86.1 3% 22.9 1-13 481 85.4 4% 12.0 1-11 481 85.4 4% 12.0 1-08 341 85.5 4% 8.5 1-06 341 85.5 4% 8.5 1-05 341 85.5 4% 8.5 1-04 378 85.6 4% 9.4 1-03 556 85.0 5% 13.9 1-02 416 84.9 4% 10.4 Totals & Averages 16698 417.5

CHAPTER 5 ENERGY ANALYSIS

Because current energy modeling software does not have the capability to model DV, a spreadsheet method was required. Energy modeling programs have no concept of vertical temperature stratification, and therefore any results from them would be unreliable. The bin method was selected because of its simplicity and the modified bin method was used because of the ability to model the building as both occupied and unoccupied. The modified bin method was created in 1983 by Knebel and allows the calculation of various building modes of operation.

Building Load Profile

Because the loads inside the building were known and the system already designed, the hand calculation of loads was not necessary. A building load profile was developed using weather data and loads from the Trane TRACE load calculation software. The full weather data can be found in appendix D. Three separate curves were then fit to these data for occupied heating, occupied cooling, and unoccupied cooling. The curves can be seen in figure 5-1. The equations fit to the data for occupied cooling, unoccupied cooling, and occupied heating can be seen in equations 5-1, 5-2, and 5-3 respectively.

= 34.710 ey 0827.0 x (5-1)

= 2.8208 ey 0226.0 x (5-2)

7.8561 xy −= 4437453 (5-3)

The occupied and unoccupied cooling data were best fit by exponential curves, while a linear curve was sufficient for the heating. This cooling curve fit results from the cooling methodology used, the total equivalent temperature differential method with time averaging

(TETD/TA) from the ASHRAE Handbook of Fundamentals. This method utilizes a transfer

function to linearize non-linear inputs. The heating methodology used was the UATD, which is an instantaneous method calculated from the U factor, the areas, and the temperature differences.

The load profiles indicate that there is no unoccupied heating as the building never reaches the heating drift-point. The balance point for heating and cooling was found at 52°F (11.1 C).

Therefore, bin analysis will be done for cooling for temperatures greater than 52°F and heating for less than 52°F.

Bin Weather Data

Bin weather data were created from the weather file of Trane TRACE. The data were grouped into two categories, occupied and unoccupied. The building was considered to be occupied from 7 AM to 6 PM, Monday through Friday. The hours unoccupied and occupied as well as the bin weather data for Gainesville, FL can be seen in table 5-1.

All Zone Level

The modified bin method determines energy consumption by the zone, system, and plant levels of the building. Because individual zone loads and the building load profile were already calculated, the zone level was completed using a summation of all zones, solely for tabulation purposes. The block space heating load and supply flow rate were fit to building load profiles.

The block load of the system was considered to be 81% of the peak load for both MV and DV.

This factor was determined by running an energy simulation on the building using TRACE. In both systems, the minimum supply flow rate was determined by the maximum of the outdoor air required and the air handler minimum of 25%. The occupied cooling zone level spreadsheets for

DV and MV can found in tables 5-2 and 5-3. The tables show that the DV system requires more supply air volume than the traditional system for the same loads. Spreadsheets for unoccupied cooling and heating can be found in the appendices.

System Level

The system level was considered as recommended by the modified bin method. The space return air dry bulb (RADB) was determined by a simple energy balance in the space and reheat was added to compensate for any over cooling due to supply air minimums. A pre-cooling coil was added to both systems in an effort to lower the chiller size and eliminate reheat. The precooling coil cools the outside air to 57°F in order to dehumidify before mixing. The mixing air dry bulb (MADB) is then lowered and further cooling or reheating is then added to the mixed air stream to reach the supply temperature. The humidity levels were tracked to determine the latent coil load which when added to the total sensible coil load determined the size of the air cooled chiller. The DV occupied system level spreadsheets (including heating and cooling) can be seen in table 5-4. This table shows the air’s properties as it moves from supply to return and is mixed with outdoor air. The DV and MV unoccupied cooling and occupied heating system level spreadsheets can be found in the appendices. The system level results were checked with energy balances to make sure that the results took all the state points of the air into account.

Plant Level

The plant level was modeled by selecting an appropriately sized air cooled chiller and using performance data from the manufacturer to determine operation at various conditions. The affect of part load operation and outdoor temperature on chiller efficiency was accounted for in the calculation. The power usage of the chiller was calculated and added to the power usage of the fan (part load operation accounted for) and the reheat used to determine the total electric consumption. This figure was then multiplied by the bin hours to determine the kilowatt hours

(kWh) used in cooling the building. It is important to note that the power consumption of the chilled water pump present in both systems was not included because there is no difference in the

MV and DV system pump use. Similarly, control system power usage was not accounted for.

The DV occupied heating and cooling plant level spreadsheets can be seen in table 5-5. The table shows how the chiller, fan, and reheat add up to the total power consumed. DV unoccupied and all MV spreadsheets are available in the appendices.

Analysis

The total energy usage of each method can be seen in table 5-6. Further, a table of checksums is presented in table 5-7.

The displacement system, if designed as specified in the ASHRAE design guidelines, would use much more energy than the mixing system due to excessive reheat caused by the dehumidification of the supply air. The ASHRAE guide does not specifically deal with the challenges of a hot and humid climate. In this climate, the higher supply temperature of the DV system adds to the reheat problem because the dehumidification temperature does not change.

The air must be cooled to 57°F (13.9 C) to remove humidity and then reheated to the supply temperature. Adding a precooling coil to the system is a way to minimize the reheat by using the same chiller to dehumidify a lower volume of outdoor air before mixing with the return air stream. The high return air temperature then keeps reheat from being necessary as it mixes with the cooled outdoor air resulting in a temperature not far from the supply temperature. Additional cooling was then done with the main cooling coil. This reheat avoidance strategy is responsible for the comparable energy use of the displacement system and mixing systems. The precooling coil was added to the mixing system as well, but less energy saving is possible using this ventilation strategy because of the lower supply and return temperatures. Both chillers were downsized because of the precooling coil eliminating the need to cool the larger volume of air.

The DV system was able to utilize a nominal 40 ton chiller to satisfy its 35 ton load while the

MV system needed a nominal 50 ton chiller to satisfy its 44 ton load.

The displacement system was designed according to the ASHRAE guidelines which should provide a draft free environment with the same or better indoor air quality as displacement ventilation; however the high supply velocities and cooling loads would be predicted to cause drafts by the REHVA guidelines. The humidity concerns should be dealt with adequately by the dehumidification of the outdoor air by the pre-cooling coil; however the energy to dehumidify reduces the energy benefit of DV found in other climates. The model predicts an energy consumption increase (penalty) of 4.6% by using DV rather than MV. This is caused by the high humidity and high supply flow rate needed to deal with the high cooling loads. DV requires more cooling energy and heating energy when occupied. The cooling energy increase is caused by an increase in fan energy and reheat that overwhelms the reduction in chiller energy. The model is in agreement with previous work that shows a reduction in chiller size and an increase in heating energy by using DV, but does not find the reduction in cooling energy that had been previously stated. Table 5-8 shows the break down of cooling energy and table 5-9 shows the energy consumption in each mode.

Figure 5-1 Building Load Profile

Table 5-1 Bin Weather Data Bin OADB (F) Occupied (hrs) Unoccupied (hrs) 0 0 0 5 0 0 10 0 0 15 0 0 20 0 0 25 0 0 30 0 0 35 0 0 40 13.8 141.2 45 54.4 337.6 50 139.6 521.4 55 200.9 407.1 60 156.3 454.8 65 299.5 766.5 70 400.6 658.4 75 372.2 1022.8 80 292.9 566.1 85 480.0 558.0 90 363.6 397.4 95 94.1 60.9 100 0 0 More 0 0 2867.9 5892.1

Table 5-2 DV Occupied Zone Level Outdoor Air Peak Block Latent Peak Block Real airflow CFM per Temp Load Load Load Airflow Airflow w/ AHU min square OADB QP QB QLA V V V foot F BTU/hr BTU/hr BTU/hr CFM CFM CFM CFM/FT2 97 215408 174481 22400 16799 13608 13608 0.91 92 142456 115389 22400 11110 8999 8999 0.602 87 94211 76311 22400 7347 5951 5951 0.398 82 62305 50467 22400 4859 3936 3936 0.263 77 41204 33375 22400 3213 2603 3402 0.227 72 27249 22072 22400 2125 1721 3402 0.227 67 18021 14597 22400 1405 1138 3402 0.227 62 11918 9653 22400 929 753 3402 0.227 57 7882 6384 22400 615 498 3402 0.227

Table 5-3 MV Occupied All Zone Level Outdoor Real Air Peak Block Latent Peak Block airflow w/ CFM per Temp Load Load Load Airflow Airflow AHU min square OADB QP QB QLA V V V foot F BTU/hr BTU/hr BTU/hr CFM CFM CFM CFM/FT2 97 2164378 286382 231969 22400 10739 10739 0.718 92 1431372 189393 153409 22400 7102 7102 0.475 87 946611 125252 101454 22400 4697 4697 0.314 82 626024 82833 67095 22400 3106 3106 0.207 77 414009 54780 44372 22400 2054 2782 0.186 72 273797 36228 29344 22400 1359 2782 0.186 67 181071 23959 19406 22400 898 2782 0.186 62 119748 15845 12834 22400 594 2782 0.186 57 79193 10479 8488 22400 393 2782 0.186

Table 5-4 DV Occupied System Level Outdoor Pre- Outdoor Mean Air Airflow Outdoor Return Cooling Mixing Air Coincident Humidity w/ Air Dry Coil Air Dry Temp Wet Bulb Ration Airflow minimum Fraction Bulb Load Bulb OADB MCWB OAW V V XOA RADB QPCC MADB F F lb/lb CFM CFM % F BTU/hr F 97 79.8 0.0181 13608 13608 21% 80.6 124736 75.6 92 78.8 0.0184 8999 8999 32% 80.6 109144 73.0 87 74.4 0.0155 5951 5951 49% 80.6 93552 69.1 82 72.4 0.0150 3936 3936 73% 80.6 77960 63.3 77 68.0 0.0126 2603 3402 85% 80.6 62368 62.9 72 62.0 0.0096 1721 3402 85% 80.6 46776 65.5 67 58.5 0.0086 1138 3402 85% 80.6 31184 67.3 62 55.2 0.0077 753 3402 85% 80.6 15592 68.4 57 50.3 0.0062 498 3402 85% 80.6 0 69.2

Table 5-4 Continued Return Mixing Cooling Outdoor Sensible Air Air Coil Reheat Coil Total Air Coil Humidity Humidity Leaving Coil Latent Coil Temp Load Ratio Ratio Humidity Load Load Load OADB QCS RAW MAW CCLAW QRHC QCL QCT F BTU/hr lb/lb lb/lb lb/lb BTU/hr BTU/hr BTU/hr 97 101021 0.00580 0.00841 0.00545 0 194793 420550 92 41930 0.00597 0.00994 0.00545 0 195510 346583 87 2851 0.00623 0.01074 0.00545 0 152188 248591 82 0 0.00663 0.01276 0.00545 22993 139207 217167 77 0 0.00682 0.01175 0.00545 31389 103712 166080 72 0 0.00682 0.00916 0.00545 33098 61116 107892 67 0 0.00682 0.00828 0.00545 34229 46640 77824 62 0 0.00682 0.00758 0.00545 34977 34977 50569 57 1765 0.00682 0.00629 0.00545 37236 13853 15618

Table 5-5 DV Occupied Plant Level Outdoor Total Total Air Coil Coil Temp Load Load Chiller Fan Reheat Total OADB Freq QCT QCT % Load EER Power Power Power Power Total F Hours BTU/hr Tons kW kW kW kW kWh 97 0 420550 35.0 99.8% 9.6 44 11.97 0 56.0 0 92 94.1 346583 28.8 82.3% 11.2 30.8 3.46 0 34.3 3227 87 363.6 248591 20.7 59.0% 12.2 20.4 1 0 21.4 7795 82 480 217167 18.1 51.6% 12.2 17.8 0.29 6.7 24.8 11903 77 292.9 166080 13.8 39.4% 11.5 14.5 0.19 9.2 23.9 6994 72 372.2 107892 8.9 25.6% 10.1 10.7 0.19 9.7 20.5 7647 67 400.6 77824 6.4 18.5% 9.2 8.4 0.19 10 18.6 7471 62 299.5 50569 4.2 12.0% 8.3 6.1 0.19 10.3 16.5 4948 57 156.3 15618 1.3 3.7% 7.0 2.2 0.19 10.9 13.3 2085 52 200.9 87918 7.3 10.6% 9.5 9.2 4.54 7.4 18.7 3748 47 139.6 66884 5.5 8.0% 8.9 7.5 5.42 9.6 19.5 2727 42 54.4 53731 4.4 6.5% 8.4 6.4 6.4 11.9 21.0 1140 37 13.8 40660 3.3 4.9% 7.9 5.1 7.49 14.2 22.3 308

Table 5-6 Energy Consumption Mode Operation kWh Occupied 69798 Unocc 42845 DV Total 112644 Occ 57999 Unocc 49665 MV Total 107664

Table 5-7 Checksums Checksums DV MV Energy kWh/ft^2 7.54 7.21 CFM/ft^2 0.911 0.799 Cooling ft^2/ton 425 360 Heating BTU/hr*ft^2 7.58 7.58

Table 5-8 DV and MV Cooling Energy Consumption Comparison DV kWh MV kWh Chiller 32879 Chiller 38003 Fan 1114 Fan 977 Reheat 27865 Reheat 16075 Tot 61857 Tot 55055

Table 5-9 DV and MV Energy Consumption Comparison Mode DV MV % change kWh kWh Occupied Cooling 61857 55054 -12.4% Occupied Heating 7941 2944 -169.7% Unoccupied Cooling 42845 49665 13.7% All Cooling 104702 104720 0.0% Total 112643 107664 -4.6%

CHAPTER 6 COST ANALYSIS

Life Cycle Cost Analysis

Life cycle cost analysis as outlined by NIST handbook 135 was used to determine the life cycle cost (LCC) of both systems. The NIST computer program BLCC5 and spreadsheets were used to determine the LCC. The parameters of the analysis are as follows,

• 20 year study period • No energy rate change forecasting/escalation • Discount rate of 3.0% from NIST • End of year discounting • Constant dollar analysis • No service, replacement, or operation costs (assumed as same for both)

Electricity Cost

The electricity rates were based on Gainesville Regional Utilities (GRU) business rates for consumption and demand. Forecasting future energy rates is extremely difficult and was not attempted. The consumptions predicted by the model were used with these rates to determine the cost of electricity. The demand for the building was determined by the operating point where the most cooling was needed and the corresponding energy needed. These values can be seen for each month in table 6-1. The rates and detailed costs can be seen in table 6-2. The model predicts that the DV system will cost $104 more in electricity cost per year, meaning that both systems have basically the same energy consumption, as DV’s advantages in chiller energy are eliminated by dehumidification and reheat.

First Costs

The system first costs were calculated by receiving estimates from contractors and suppliers. The following companies provided information for the analysis,

• Mechanical Contractors Inc – Duct, VAV boxes, Mixing diffusers, Return air grills, Labor data, Air balancing, Chases • Price Air Distribution – Displacement diffusers • Grainger Industrial Supply – Baseboard heaters • Trane – Chillers, Air Handlers

Some equipment normally included in LCC analysis of HVAC systems was not included, such as,

• Fire dampers • Control dampers, panels • Chilled water system (pump, piping) • Replacement, service, operation (same for both systems)

Even though the chillers are different sizes, there was assumed not to be a substantial difference in the cost of the piping systems due to the small size of the chillers. It should be noted that this will not be the case in larger systems.

First cost data input data used for the analysis can be found in table 6-3. Table 6-4 shows the calculation procedure for the first cost for both systems, as well as financial checksums. All costs include labor. The DV system costs $187,000 compared to $154,000. The higher costs are a result of a larger air handler, more duct (for the higher flowrate and drops to the floor level, more expensive diffusers (almost triple the cost of normal diffusers), and especially the chases need to hide the duct drops to the floor level. The drops range in size from 10 inches (25.4 cm) round to 18”x8” (45.7x20.3 cm). A normal wall cannot accommodate anything larger than 3”

(7.6 cm) so chases are needed. The DV diffuser details, including cost, face area, and drop size, can be found in table 6-5. This study does not consider the costs of the lost floor space for these chases and diffusers.

The LCC was conducted using both the BLCC5 program and a spreadsheet analysis. The inputs were the system first cost and energy usage only. The discount rate used was given by the

US Office of Management and Budget document “OMB Circular No. A-94, Appendix C” as

instructed by the NIST Handbook 135. The real discount rate for a 20 year study period as of

January 2007 is 3.0%. Table 6-6 shows the yearly present value costs of the MV and DV system and totals them to find the LCC. These results are in agreement with the BLCC5 program.

The LCC of the DV system is $417,342 compared to $382,715. This is mainly a result of the substantially higher first cost which is never paid back due to a lack of energy savings. The energy penalty paid by the DV system has a very small impact. It is impossible to recommend this system on a LCC basis.

Table 6-1 Monthly Electric Demand Month Demand (kW) $ DV MV DV MV Jan 18.9 24.4 $170 $220 Feb 22.4 22.5 $202 $202 Mar 22.4 22.5 $202 $202 Apr 21.4 25.9 $192 $233 May 21.4 25.9 $192 $233 Jun 34.2 39.3 $308 $353 Jul 34.2 39.3 $308 $353 Aug 55.9 61.2 $503 $551 Sep 34.2 39.3 $308 $353 Oct 21.4 25.9 $192 $233 Nov 22.4 22.5 $202 $202 Dec 17.0 26.2 $153 $236 Tot $2,939 $3,378

Table 6-2 Electric Rates and Costs GRU DV MV Consumption Charge Service $16 Customer Charge $192 $192 per kWh for first $0.06 1500 $1,116 $1,116 $0.08 per kWh over 1500 $7,571 $7,173 Demand $33 Customer Charge $396 $396 $9 per kW demand $2,939 $3,378 $0.03 per kWh $3,266 $3,122 Totals $15,482 $15,377 Savings ($104)

Table 6-3 First Cost Unit Inputs DV MV Unit 13608 11933 CFM 40 50 Tons supply 47 73 grills VAV 31 31 boxes 29 26 heaters 63 63 RA grills

Table 6-4 System First Costs Category Item Rate Units DV MV Units Duct Standard 0.6 lbs/cfm 8165 7160 lbs 4 $/lb $32,658 $28,638 $ Drops for DV 75 $/drop $3,525 $ - $ Grills 2x2 supply 100 $/unit $ - $7,300 $ Return 35 $/unit $2,205 $2,205 $ Air Balancing 25 $/unit $1,175 $1,825 $ DV from Grills supplier $20,281 $ - $ fro DV Chases grills 400 $/grill $18,800 $ - $ VAVs W/ heat 1500 $/unit $ - $ - $ No heat 1000 $/unit $31,000 $31,000 $ AHUs Trane 2 $/cfm $27,215 $23,865 $ Chillers Trane 1000 $/ton $40,000 $50,000 $ BB 1 kW Heaters heaters 350 $/unit $10,150 $9,100 $ Totals $187,009 $153,933 $ $/ton $4,675 $3,078 $ $/ft2 $12.50 $10.29 $

Table 6-5 DV Diffuser Data Max V Face @ 40 # Drop Areas fpm needed Size Cost ft2 CFM in $ 7.7 310 21 10 $6,756 11.6 463 17 10 $8,152 13.5 540 5 14x6 $2,764 19.1 763 2 18x8 $1,519 15.8 633 2 18x8 $1,090 47 $20,281

Table 6-6 Life Cycle Cost DV MV Year Costs Electricity Costs Electricity 0 $187,009 $ - $153,933 $ - 1 $ - $15,031 $ - $14,929 2 $ - $14,593 $ - $14,494 3 $ - $14,168 $ - $14,072 4 $ - $13,755 $ - $13,662 5 $ - $13,354 $ - $13,264 6 $ - $12,965 $ - $12,878 7 $ - $12,588 $ - $12,503 8 $ - $12,221 $ - $12,139 9 $ - $11,865 $ - $11,785 10 $ - $11,520 $ - $11,442 11 $ - $11,184 $ - $11,109 12 $ - $10,858 $ - $10,785 13 $ - $10,542 $ - $10,471 14 $ - $10,235 $ - $10,166 15 $ - $9,937 $ - $9,870 16 $ - $9,647 $ - $9,582 17 $ - $9,366 $ - $9,303 18 $ - $9,094 $ - $9,032 19 $ - $8,829 $ - $8,769 20 $ - $8,572 $ - $8,514 Total (present worth) $417,342 $382,715

CHAPTER 7 CONCLUSIONS

The survey predicts that a DV system will consume 4.6% more electricity than a traditional

MV system when implemented in an office building in Gainesville, Florida. This penalty is due to an increase in cooling energy from the fan and reheat coil which eliminate the savings from chiller energy. Heating energy using DV was also much greater. The life cycle cost of the DV system was predicted to be $34,600 more than a traditional MV system, about 8.3% more. The first cost premium is never paid back as there is no energy savings.

Design

The DV design process as proposed by ASHRAE has weaknesses in several areas. Many of the equations used to calculate the design parameters are noted to not be accurate, but the process is meant to be simplified by this procedure. The cooling load limits and diffuser face velocities are much higher than other published values, which could lead to an increase in drafts and high temperature gradient risk. The humidity in the space is not given adequate treatment to confidently use the process in a hot and humid climate like Florida. Even so, the guide is the only process meant for use in high cooling load areas like the US and previous research has shown that at worst, DV will function as a mixing system. For these reasons, it must be assumed that the system described in this study will maintain indoor air quality only equal to a normal mixing system, due to the large amount of return air cycling and may have the potential for draft issues.

Energy Analysis

The energy analysis portion of this study utilized the modified bin method, and thus is subject to the weaknesses of this method. All bin methods using dry bulb temperature and mean coincident wet bulb temperature have been considered to underestimate the latent cooling loads

in humid climates. This may be a weakness of this study as well, but the effect should be felt equally on the DV and MV systems. Also, bin methods are complicated and are not easily adapted to some control strategies that could save energy such as supply air temperature rest, fan pressure reset, morning warmup, etc. These results should serve as a comparison only for the simple systems described and serve as a starting point for considering more features. Until more actual data on real system performance are available and current energy modeling programs expand to include displacement ventilation, more accurate results with more complicated systems are unlikely to be created.

Life Cycle Cost Analysis

The LCC provides an estimation of what the real ownership cost will be of a displacement versus a mixing system. Because this LCC only has two components, first cost and energy cost, the LCC is extremely dependent on the accuracy of the electricity cost and the discount rate.

Energy cost changes are extremely hard to predict, but most assume they will continue to rise in the future, so any difference in LCC may be magnified with time.

Recommendations for Future Research

Given the weaknesses previously mentioned, more research is needed. Specifically the following should be considered,

• Actual experimental data are needed to demonstrate the ability of DV to meet high cooling loads and provide a comfortable environment in terms of humidity. A pilot project using systems similar to the ones described is needed. • The effects of cross ventilation and occupant movement need to be studied • The expansion of energy simulation programs to include DV would similarly serve as a huge step forward. • Expanding the current spreadsheet model to include other control systems • Further study on contaminant distribution seeking to determine the location of contaminants using DV • Further study on the humidity levels inside the space, specifically where stratification occurs.

Several studies predict that DV can save energy and but this study does not repeat those findings for this climate. At this point it cannot be recommended to implement DV without other energy saving strategies not considered in this study. Further, the lack of real buildings using the technology and the admitted weaknesses of the design guides necessitate more research before DV can become a usable technology.

APPENDIX A BUILDING FLOORPLAN

Figure A-1 Building Floorplan

APPENDIX B DISPLACEMENT VENTILATION SPREADSHEETS

Design

Table B-1 Input Parameters for DV Design aoe 0.295 ∆Thf (F) 3.6 H (ft) 12 TSET, T-head (F) 75 al 0.132 ρ (lb/ft3) 0.0749 αr 1 T-foot (F) 71.4 aex 0.185 Cp (BTU/lb*F) 0.24 αcf 1 60*ρ*Cp 1.08

Table B-2 Output of DV Design SADB 68.7 F Max OA % 17% % OA CFM 2887 CFM Avg RADB 83.4 F

Table B-3 DV Design Sizing Load # Area Density Zone Name Rooms A QOE QL QEX Q Total Q/A Q/A ft2 BTU/hr BTU/hr BTU/hr BTU/hr BTU/hr*ft2 W/m2 3 2 894 6498 4638 0 11136 12.5 39.3 2 2 213 3242 1105 0 4347 20.4 64.4 16 2 644 6485 3341 0 9826 15.3 48.1 12 3 723 2695 3751 0 6446 8.9 28.1 5 2 258 1797 1338 0 3135 12.2 38.3 1 2 525 1148 2724 0 3872 7.4 23.3 4 2 308 2121 1598 0 3719 12.1 38.1 9 2 252 1797 1307 0 3104 12.3 38.9 6 2 227 1797 1178 0 2975 13.1 41.3 7 2 257 1797 1333 0 3130 12.2 38.4 10 2 194 898 1006 0 1905 9.8 31.0 8 2 405 1797 1006 0 2803 6.9 21.8 14 3 659 0 3419 0 3419 5.2 16.4 13 2 300 6660 1556 0 8217 27.4 86.4 17 2 285 500 1401 2440 4341 15.2 48.0 21 1 582 4492 2860 2177 9530 16.4 51.7 22 1 640 3594 3145 3019 9758 15.2 48.1 23 1 962 2047 4728 9939 16714 17.4 54.8 24 1 546 400 2385 7247 10032 18.4 58.0 25 1 831 1797 4311 20248 26356 31.7 100.1 19 1 1315 6000 6463 6436 18899 14.4 45.3 20 1 515 3898 2531 2286 8716 16.9 53.4 18 4 641 2141 2800 9922 14863 23.2 73.1 1-14 3 525 1863 2259 13118 17239 32.8 103.6 1-13 2 308 1242 1346 6106 8693 28.2 89.0 1-11 2 308 1242 1346 6106 8693 28.2 89.0 1-8 1 227 696 992 4499 6187 27.3 86.0 1-6 1 227 696 992 4499 6187 27.3 86.0 1-5 1 227 696 992 4499 6187 27.3 86.0 1-4 1 255 696 1114 5061 6871 26.9 85.0 1-3 1 384 1648 1678 6449 9775 25.5 80.3 1-1 1 306 1148 1474 4636 7259 23.7 74.8 Totals & Averages 14,943 73,532 72,116 118,687 264,335

Table B-3 Continued Cooling V Zone Flow VC Effective OA Flow Name n n Min Rate VC Min ness η Rate V62 XOA VOA ACPH ACPH CFM CFM CFM CFM 3 3.64 3.64 651 651 1.63 96 15% 59 2 6.66 6.66 284 284 2.44 18 6% 7 16 4.71 4.71 606 606 1.98 44 7% 22 12 2.30 2.30 332 332 1.05 58 18% 56 5 3.53 3.53 182 182 1.59 25 14% 16 1 1.71 1.71 180 180 0.75 42 23% 55 4 3.50 3.50 215 215 1.58 17 8% 11 9 3.59 3.59 181 181 1.61 25 14% 16 6 3.89 3.89 177 177 1.72 24 13% 14 7 3.54 3.54 182 182 1.59 25 14% 16 10 2.64 2.64 102 102 1.21 12 12% 10 8 2.11 2.11 171 171 1.19 29 17% 25 14 0.88 0.88 116 116 0.30 16 14% 53 13 9.32 9.32 559 559 2.79 8 1% 3 17 3.54 1.50 202 86 1.12 5 3% 5 21 4.66 3.77 542 439 1.75 60 11% 34 22 4.09 2.97 524 380 1.51 58 11% 39 23 4.11 1.64 790 316 1.24 130 17% 105 24 4.18 1.02 457 111 1.16 76 17% 65 25 7.51 1.70 1248 283 1.55 110 9% 71 19 3.73 2.57 982 676 1.38 104 11% 75 20 4.77 3.71 491 382 1.74 88 18% 50 18 5.70 2.01 731 258 1.50 34 5% 23 1-14 8.03 2.08 843 218 1.65 47 6% 28 1-13 7.00 2.27 431 140 1.62 28 7% 18 1-11 7.00 2.27 431 140 1.62 28 7% 18 1-8 6.63 1.91 301 87 1.55 19 6% 12 1-6 6.63 1.91 301 87 1.55 19 6% 12 1-5 6.63 1.91 301 87 1.55 19 6% 12 1-4 6.51 1.78 332 91 1.52 20 6% 13 1-3 6.37 2.37 490 182 1.6 43 9% 27 1-1 5.85 2.24 358 137 1.53 23 7% 15 Totals & Averages 13,693 8,038 1,351 985

Table B-3 Continued SADB- VOA V Min Zone Name V θf SADBi RADB SADBi Min Min % CFM F F F CFM CFM 3 651 0.388 65.2 81.1 3.4 555 651 78% 2 284 0.258 67.7 81.9 0.9 102 283 93% 16 606 0.329 66.4 81.5 2.2 252 606 85% 12 332 0.502 62.3 80.4 6.3 337 337 66% 5 182 0.396 65.0 81.0 3.6 147 182 77% 1 180 0.575 59.9 79.9 8.8 240 240 75% 4 215 0.398 65.0 81.0 3.7 99 215 77% 9 181 0.392 65.1 81.1 3.5 145 181 78% 6 177 0.373 65.5 81.2 3.1 136 176 80% 7 182 0.395 65.0 81.0 3.6 147 181 77% 10 102 0.467 63.3 80.6 5.3 71 102 69% 8 171 0.523 63.4 78.7 5.3 169 170 65% 14 116 0.724 51.6 78.9 17 91 116 38% 13 559 0.199 68.6 82.3 0.0 47 558 100% 17 202 0.395 63.5 83.5 5.2 31 85 31% 21 542 0.332 65.9 82.3 2.7 346 438 67% 22 524 0.361 65.1 82.4 3.5 337 380 58% 23 790 0.360 64.3 83.9 4.4 754 754 74% 24 457 0.356 64.1 84.5 4.5 436 436 74% 25 1248 0.235 66.7 86.4 1.9 634 634 46% 19 982 0.382 64.5 82.4 4.1 600 675 53% 20 491 0.327 66.0 82.5 2.7 507 507 87% 18 731 0.289 65.9 84.8 2.7 197 257 30% 1-14 843 0.223 67.1 86.1 1.5 268 268 29% 1-13 431 0.248 66.7 85.4 1.9 164 164 34% 1-11 431 0.248 66.7 85.4 1.9 164 164 34% 1-8 301 0.259 66.4 85.5 2.2 107 107 32% 1-6 301 0.259 66.4 85.5 2.2 107 107 32% 1-5 301 0.259 66.4 85.5 2.2 107 107 32% 1-4 332 0.262 66.3 85.6 2.3 117 117 31% 1-3 490 0.266 66.4 85.0 2.2 248 248 45% 1-1 358 0.283 66.0 84.9 2.6 135 137 33% Totals & Averages 13,693 7813 9600

Table B-3 Continued V ∆Thf Tf Th θf RADB Diffuser Zone Name recalc recalc recalc recalc recalc Recalc XOA RADB*V Area A CFM F F C F % CFM*F ft2 3 832 2.8 72.8 75.6 0.3 81.1 7% 67503 20.8 2 304 3.4 71.9 75.3 0.2 81.9 2% 24926 7.6 16 713 3.1 72.5 75.5 0.3 81.5 3% 58055 17.8 12 512 2.3 73.3 75.6 0.4 80.4 11% 41153 12.8 5 235 2.8 72.8 75.6 0.3 81.0 7% 19070 5.9 1 321 2.0 73.5 75.5 0.4 79.9 17% 25610 8.0 4 279 2.8 72.9 75.6 0.3 81.0 4% 22640 7.0 9 232 2.8 72.8 75.6 0.3 81.1 7% 18845 5.8 6 221 2.9 72.7 75.6 0.3 81.2 6% 17909 5.5 7 235 2.8 72.8 75.6 0.3 81.0 7% 19033 5.9 10 148 2.5 73.2 75.7 0.4 80.6 7% 11965 3.7 8 261 2.4 72.9 75.2 0.4 78.7 9% 20512 6.5 14 310 1.4 73.8 75.1 0.5 78.9 17% 24448 7.7 13 559 3.6 71.4 75.0 0.2 82.3 1% 46009 14.0 17 273 2.7 73.5 76.2 0.3 83.5 2% 22746 6.8 21 650 3.0 72.7 75.7 0.3 82.3 5% 53473 16.2 22 658 2.9 73.0 75.8 0.3 82.4 6% 54266 16.5 23 1016 2.8 73.3 76.1 0.3 83.9 10% 85267 25.4 24 588 2.8 73.4 76.2 0.3 84.5 11% 49690 14.7 25 1382 3.2 72.5 75.8 0.2 86.4 5% 119371 34.6 19 1277 2.8 73.1 75.9 0.3 82.4 6% 105222 31.9 20 586 3.0 72.7 75.7 0.3 82.5 9% 48339 14.7 18 855 3.1 72.8 75.9 0.3 84.8 3% 72481 21.4 1-14 917 3.3 72.3 75.6 0.2 86.1 3% 78991 22.9 1-13 481 3.2 72.5 75.7 0.2 85.4 4% 41087 12.0 1-11 481 3.2 72.5 75.7 0.2 85.4 4% 41087 12.0 1-8 341 3.2 72.7 75.8 0.2 85.5 4% 29135 8.5 1-6 341 3.2 72.7 75.8 0.2 85.5 4% 29135 8.5 1-5 341 3.2 72.7 75.8 0.2 85.5 4% 29135 8.5 1-4 378 3.2 72.7 75.9 0.2 85.6 4% 32310 9.4 1-3 556 3.2 72.6 75.8 0.2 85.0 5% 47271 13.9 1-1 416 3.1 72.8 75.9 0.3 84.9 4% 35314 10.4 Totals & Averages 16698 6% 1391994 417.5

Zone

Table B-4 DV Occupied Zone Inputs Altitude 155 ft Peak V 16698 CFM Block Load 215408 BTU/hr Area 14,943 ft^2 Min V 2887 CFM Block V 13608 CFM ∑Q 264,335 BTU/hr SADB 68.7 F AHU Size 13608 CFM Eqn Coeff 710.34 a TSET 75 F AHU Min 25% % Y=ae^bx 0.0827 b # People 112 AHU Min 3401.88 CFM

Table B-5 DV Occupied All Zones Q Q Peak Block Q Zone Block V w/ OADB Zones Zones Latent Peak V V min V/A F BTU/hr BTU/hr BTU/hr CFM CFM CFM CFM/ft2 97 215408 174481 22400 16799 13608 13608 0.910 92 142456 115389 22400 11110 8999 8999 0.602 87 94211 76311 22400 7347 5951 5951 0.398 82 62305 50467 22400 4859 3936 3936 0.263 77 41204 33375 22400 3213 2603 3402 0.227 72 27249 22072 22400 2125 1721 3402 0.227 67 18021 14597 22400 1405 1138 3402 0.227 62 11918 9653 22400 929 753 3402 0.227 57 7882 6384 22400 615 498 3402 0.227

Table B-6 DV Unoccupied Zone Inputs Altitude 155 ft Peak V 6664 CFM Block Load 124,561 BTU/hr Area 14943 ft^2 Min V 0 CFM Block V 6664 CFM ∑Q 124,561 BTU/hr SADB 59 F AHU Size 13608 CFM Eqn Coeff 8208.2 a TSET 68 F AHU Min 0.25 % Y=ae^bx 0.0226 b # People 0 AHU Min 3402 CFM

Table B-7 DV Unoccupied All Zones Q Block Q peak block Q Zone V w/ OADB zones zones Latent Peak V min V/A F BTU/hr BTU/hr BTU/hr CFM CFM CFM/ft2 97 124561 101505 0 6664 5431 0.363 92 111251 90659 0 5952 4851 0.324 87 99364 80972 0 5316 4332 0.289 82 88747 72321 0 4748 3869 0.258 77 79265 64593 0 4241 3456 0.231 72 70795 57691 0 3788 3402 0.227 67 63231 51527 0 3383 3402 0.227 62 56475 46021 0 3022 3402 0.227 57 50440 41104 0 2699 3402 0.227

System

Table B-8 DV Occupied Cooling System Inputs Altitude 155 ft V Peak 16698 CFM PCLADB 57 F CCLRH 55% Area 14,943 ft^2 RADB 80.6 F CCLAW 38.2 gr/lb SADB 68.7 CCLADB 57 F VOA 2887 CFM CCLAW 0.00545 lb/lb TSET 75

Table B-9 DV Occupied Cooling System Level Outdoor Outdoor Outdoor Return Air Outdoor Mean Air Air Air Real Dry Bulb Air Dry Coincident Humidity Humidity Relative w. % (without QRH from Bulb Wet Bulb Ratio Ratio Humidity Total min OA reheat) AHU min OADB MCWB OAW OAW OARH V V XOA RADB F F gr/lb lb/lb % CFM CFM % F BTU/hr 97 79.8 126 0.0181 48 13608 13608 21% 80.6 0 92 78.8 128 0.0184 56 8999 8999 32% 80.6 0 87 74.4 108 0.0155 56 5951 5951 49% 80.6 0 82 72.4 104 0.0150 64 3936 3936 73% 80.6 0 77 68.0 88 0.0126 63 2603 3402 85% 77.8 10245 72 62.0 67 0.0096 57 1721 3402 85% 74.7 21548 67 58.5 59 0.0086 61 1138 3402 85% 72.7 29023 62 55.2 54 0.0077 65 753 3402 85% 71.3 33967 57 50.3 43 0.0062 63 498 3402 85% 70.4 37236

Table B-9 Continued Cooling Cooling Return Outdoor Pre Pre Mix Air Coil Coil Preheat Air Air Dry Real Cooling Cooling Dry Leaving Entering Leaving Coil Relative Bulb RADB Sensible Latent Bulb Dry Bulb Air DB Dry Bulb Load Humidity OADB RADB QPCC QPCC MADB CCLADB CCEADB PHLADB QCS RARH F F BTU/hr BTU/hr F F F F BTU/hr % 97 80.6 124736 178923 75.6 57 75.6 75.6 101021 26 92 80.6 109144 182017 73.0 57 73.0 73.0 41930 27 87 80.6 93552 142129 69.1 57 69.1 69.1 2851 28 82 80.6 77960 134613 63.3 57 63.3 63.3 0 30 77 80.6 62368 101494 60.6 57 60.6 60.6 0 31 72 80.6 46776 58669 60.6 57 60.6 60.6 0 31 67 80.6 31184 44081 60.6 57 60.6 60.6 0 31 62 80.6 15592 32321 60.6 57 60.6 60.6 0 31 57 80.6 0 11073 60.6 57 60.6 60.6 0 31

Table B-9 Continued Cooling Return Return Mixing Mixing Mixing Coil Outdoor Air Air Air Air Air Leaving Air Dry Humidity Humidity Humidity Humidity Relative Humidity Re Latent Total Bulb Ratio Ratio Ratio Ratio Humidity Ratio Heat Load Load OADB RAW RAW MAW MAW MARH CCLAW QRHC QCL QCT F lb/lb gr/lb lb/lb gr/lb % lb/lb BTU/hr BTU/hr BTU/hr 97 0.00580 40 0.00572 40 30 0.00546 0 196570 422327 92 0.00597 41 0.00581 40 33 0.00546 0 197229 348303 87 0.00623 43 0.00586 41 39 0.00546 0 153661 250064 82 0.00663 46 0.00577 40 46 0.00546 22993 140580 218540 77 0.00682 47 0.00566 39 50 0.00546 40085 104882 167250 72 0.00682 47 0.00566 39 50 0.00546 51388 62057 108832 67 0.00682 47 0.00566 39 50 0.00546 58863 47468 78652 62 0.00682 47 0.00566 39 50 0.00546 63806 35708 51300 57 0.00682 47 0.00566 39 50 0.00546 67076 14461 14461

Table B-10 DV Unoccupied Cooling System Level Return Outdoor Outdoor Outdoor Air Dry QRH Outdoor Mean Air Air Air Real Bulb from Air Dry Coincident Humidity Humidity Relative w. Outdoor (without AHU Bulb Wet Bulb Ratio Ratio Humidity Total min Air reheat) min OADB MCWB OAW OAW OARH V V XOA RADB F F gr/lb lb/lb % CFM CFM % F BTU/hr 97 79.8 126 0.0181 47.8 5431 5431 0% 76.3 0 92 78.8 128 0.0184 56.4 4851 4851 0% 76.3 0 87 74.4 108 0.0155 56.0 4332 4332 0% 76.3 0 82 72.4 104 0.015 63.6 3869 3869 0% 76.3 0 77 67.9 88 0.0126 63.4 3456 3456 0% 76.3 0 72 61.9 67 0.0096 57.1 3402 3402 0% 74.7 5892 67 58.5 59 0.0086 60.6 3402 3402 0% 73.0 12056 62 55.1 54 0.0077 65.2 3402 3402 0% 71.5 17562 57 50.2 43 0.0062 62.8 3402 3402 0% 70.2 22479

Table B-10 Continued Cooling Cooling Return Outdoor Pre Mix Air Coil Coil Preheat Air Air Dry Real Cooling Dry Leaving Entering Leaving Pre Coil Relative Bulb RADB Sensible Bulb Dry Bulb Air DB Dry Bulb Heat Load Humidity OADB RADB QPCC MADB CCLADB CCEADB PHLADB QPHC QCS RARH F F BTU/hr F F F F BTU/hr BTU/hr % 97 76 0 76.3 57 76.3 76.3 0 113325 28 92 76 0 76.3 57 76.3 76.3 0 101216 28 87 76 0 76.3 57 76.3 76.3 0 90401 28 82 76 0 76.3 57 76.3 76.3 0 80742 28 77 76 0 76.3 57 76.3 76.3 0 72115 28 72 76 0 76.3 57 76.3 76.3 0 70987 30 67 76 0 76.3 57 76.3 76.3 0 70987 31 62 76 0 76.3 57 76.3 76.3 0 70987 33 57 76 0 76.3 57 76.3 76.3 0 70987 35

Table B-10 Continued Cooling Return Return Mixing Mixing Mixing Coil Outdoor Air Air Air Air Air Leaving Air Dry Humidity Humidity Humidity Humidity Relative Humidity Re Latent Total Bulb Ratio Ratio Ratio Ratio Humidity Ratio Heat Load Load OADB RAW RAW MAW MAW MARH CCLAW QRHC QCL QCT F lb/lb gr/lb lb/lb gr/lb % lb/lb BTU/hr BTU/hr BTU/hr 97 0.00546 38 0.00545 38 28 0.005457 11820 0 113325 92 0.00546 38 0.00545 38 28 0.005457 10557 0 101216 87 0.00546 38 0.00545 38 28 0.005457 9429 0 90401 82 0.00546 38 0.00545 38 28 0.005457 8421 0 80742 77 0.00546 38 0.00545 38 28 0.005457 7521 0 72115 72 0.00546 38 0.00545 38 30 0.005457 7404 0 70987 67 0.00546 38 0.00545 38 32 0.005457 7404 0 70987 62 0.00546 38 0.00545 38 33 0.005457 7404 0 70987 57 0.00546 38 0.00545 38 35 0.005457 7404 0 70987

Plant

Figure B-1 Part Load Operation Data for 40 ton Trane Chiller

Figure B-2 Effect of Ambient Temperature on 40 ton Trane Chiller

Table B-11 DV Occupied Cooling Plant Level Coil Coil Total Total Load Load % Real OADB Freq QCT QCT Load EER Factor EER Factor EER F Hours BTU/hr Tons Load Temp 97 0 422327 35.1 100% 9.7 1.0 9.7 1.0 9.5 92 94.1 348303 29.0 82% 10.5 1.1 10.4 1.1 11.2 87 363.6 250064 20.8 59% 10.6 1.1 11.2 1.1 12.1 82 480.0 218540 18.2 52% 10.4 1.1 11.5 1.2 12.2 77 292.9 167250 13.9 40% 9.8 1.0 11.5 1.2 11.5 72 372.2 108832 9.0 26% 8.6 0.9 11.5 1.2 10.1 67 400.6 78652 6.5 19% 7.9 0.8 11.5 1.2 9.2 62 299.5 51300 4.2 12% 7.1 0.7 11.5 1.2 8.3 57 156.3 14461 1.2 3% 5.9 0.6 11.5 1.2 6.9

Table B-11 Continued OADB Chiller Fan Reheat Total F kW kWh SP in kW kWh BTU/hr kW kWh kW kWh 97 44.3 0 4.00 11.97 0 0 0 0 56.3 0 92 31.1 2922 1.75 3.46 326 0 0 0 34.5 3248 87 20.6 7489 0.77 1.00 364 0 0 0 21.6 7853 82 17.9 8597 0.33 0.29 139 22993 6.7 3235 24.9 11971 77 14.6 4279 0.25 0.19 55 40085 11.7 3441 26.5 7775 72 10.8 4003 0.25 0.19 70 51388 15.1 5606 26.0 9679 67 8.5 3412 0.25 0.19 75 58863 17.3 6911 26.0 10399 62 6.2 1848 0.25 0.19 56 63806 18.7 5600 25.1 7504 57 2.1 327 0.25 0.19 29 67076 19.7 3072 21.9 3428

Table B-12 DV Unoccupied Cooling Plant Level Coil Coil Total Total Load Load OADB Freq QCT QCT % Load EER Factor EER Factor Real EER F Hours BTU/hr Tons Load Temp 97 0 113325 9.4 26.9% 8.8 0.9 9.7 1.0 8.6 92 60.9 101216 8.4 24.0% 8.5 0.9 10.4 1.1 9.0 87 397.4 90401 7.5 21.5% 8.2 0.8 11.2 1.1 9.4 82 558.0 80742 6.7 19.2% 7.9 0.8 12.0 1.2 9.7 77 566.1 72115 6.0 17.1% 7.7 0.8 12.7 1.3 10.0 72 1022.8 70987 5.9 16.9% 7.7 0.8 13.5 1.4 10.6 67 658.4 70987 5.9 16.9% 7.7 0.8 14.2 1.5 11.2 62 766.5 70987 5.9 16.9% 7.7 0.8 15.0 1.5 11.8 57 454.8 70987 5.9 16.9% 7.7 0.8 15.8 1.6 12.3

Table B-12 Continued OADB Chiller Fan Reheat Total F kW kWh SP in kW kWh BTU/hr kW kWh Kw kWh 97 13.1 0 0.64 0.8 0 11820 3.5 0 17.3 0 92 11.2 683 0.51 0.5 33 10557 3.1 188 14.9 905 87 9.6 3835 0.41 0.4 154 9429 2.8 1098 12.8 5087 82 8.3 4647 0.32 0.3 154 8421 2.5 1377 11.1 6178 77 7.2 4080 0.26 0.2 111 7521 2.2 1248 9.6 5439 72 6.7 6875 0.25 0.2 191 7404 2.2 2219 9.1 9286 67 6.4 4189 0.25 0.2 123 7404 2.2 1429 8.7 5741 62 6.0 4631 0.25 0.2 143 7404 2.2 1663 8.4 6437 57 5.7 2615 0.25 0.2 85 7404 2.2 987 8.1 3686

Heating

Table B-13 DV Occupied Heating Design and Zone Level Zone Building QBB*1.1 Name Area A QOE QL Heat Loss QBB oversizing QBB QOE+QBB Q Total ft2 BTU/hr BTU/hr Btu/hr-F BTU/hr BTU/hr kW BTU/hr BTU/hr 3 894 3898 4638 0 0 0 0 3898 8536 2 213 3242 1105 0 0 0 0 3242 4347 16 644 6485 3341 0 0 0 0 6485 9826 12 723 2695 3751 0 0 0 0 2695 6446 5 258 1797 1338 0 0 0 0 1797 3135 1 525 1148 2724 0 0 0 0 1148 3872 4 308 2121 1598 0 0 0 0 2121 3719 9 252 1797 1307 0 0 0 0 1797 3104 6 227 1797 1178 0 0 0 0 1797 2975 7 257 1797 1333 0 0 0 0 1797 3130 10 194 898 1006 0 0 0 0 898 1905 8 405 1797 2101 0 0 0 0 1797 3898 14 659 0 3419 0 0 0 0 0 3419 13 300 6660 1556 0 0 0 0 6660 8217 17 285 500 1479 54.43 1959 2155 0.632 2655 4134 21 582 4492 3019 49.48 1781 1959 0.574 6452 9471 22 640 3594 3320 68.83 2478 2726 0.799 6320 9640 23 962 2047 4991 57.16 2058 2264 0.663 4311 9301 24 546 400 2683 54.95 1978 2176 0.638 2576 5260 25 831 1797 4311 170.4 6134 6748 1.978 8545 12856 19 1315 6000 6822 113.24 4077 4484 1.314 10484 17306 20 515 3898 2672 51.95 1870 2057 0.603 5956 8627 18 641 2141 2882 120.37 4333 4767 1.397 6907 9790 1-14 525 1863 2259 170.96 6155 6770 1.984 8633 10892 1-13 308 1242 1346 62.67 2256 2482 0.727 3724 5070 1-11 308 1242 1346 62.67 2256 2482 0.727 3724 5070 1-8 227 696 992 46.18 1662 1829 0.536 2525 3516 1-6 227 696 992 46.18 1662 1829 0.536 2525 3516 1-5 227 696 992 46.18 1662 1829 0.536 2525 3516 1-4 255 696 1114 51.95 1870 2057 0.603 2753 3867 1-3 384 1648 1992 118.45 4264 4691 1.375 6339 8331 1-1 306 1148 1556 79.16 2850 3135 0.919 4283 5840 Totals & Averages 14,943 1,425 51,308 56,439 202,532

Table B-14 DV Heating System Level Heat Internal OADB Loss Load QBB V RADB MADB QCS QCS F BTU/hr BTU/hr BTU/hr CFM F F BTU/hr Tons 52 22804 145,648 25084 9849 73.9 67.5 61619 5.1 47 29930 145,648 32923 10445 74.0 66.5 54530 4.5 42 37056 145,648 40762 11040 74.1 65.7 47544 3.9 37 44182 145,648 48600 11636 74.1 64.9 40660 3.4 32 51308 145,648 56439 12231 74.2 64.2 33878 2.8

Table B-14 Continued OADB OAW OAW OARH QL % OA RARH RAW RAW F gr/lb lb/lb % BTU/hr % lb/lb gr/lb 52 43.7 0.00624 62 22400 29% 33 0.00592 41.5 47 36.2 0.00518 63 22400 28% 33 0.0059 41.3 42 33.0 0.00471 69 22400 26% 33 0.00587 41.1 37 22.1 0.00316 56 22400 25% 32 0.00585 41.0 32 22.1 0.00316 56 22400 24% 32 0.00583 40.8

Table B-14 Continued OADB MAW MAW MARH CCLAW QRHC QCL QCT F lb/lb gr/lb % lb/lb BTU/hr BTU/hr BTU/hr 52 0.00602 42.1 42 0.00545 0 26907 88526 47 0.00570 39.9 41 0.00545 0 12354 66884 42 0.00557 39.0 41 0.00545 0 6187 53731 37 0.00518 36.3 39 0.00518 0 0 40660 32 0.00520 36.4 41 0.00520 0 0 33878

Table B-15 DV Heating Plant Level Coil Coil Total Total Load Load % Real OADB Freq QCT QCT Load EER Factor EER Factor EER F Hours BTU/hr Tons Load Temp 52 200.9 88526 7.4 10.6% 8.1 0.8 11.5 1.2 9.5 47 139.6 66884 5.6 8.0% 7.5 0.8 11.5 1.2 8.9 42 54.4 53731 4.5 6.5% 7.2 0.7 11.5 1.2 8.4 37 13.8 40660 3.4 4.9% 6.8 0.7 11.5 1.2 7.9

Table B-15 Continued OADB Chiller Fan Reheat Total F kW kWh SP in kW kWh BTU/hr kW kWh kW kWh 52 9.3 1864 2.1 4.5 912 25084 7.4 1477 18.7 3762 47 7.6 1054 2.36 5.4 756 32923 9.6 1347 19.6 2730 42 6.4 348 2.63 6.4 348 40762 11.9 650 21.0 1141 37 5.1 71 2.92 7.5 104 48600 14.2 197 22.3 309

APPENDIX C MIXING VENTILATION SPREADSHEETS

Design

Table C-1 MV Design Level Zone Name Area QT V Load Density V62.1 XOA ft^2 BTU/hr CFM BTU/hr*ft2 CFM 3 894 11136 573 9.5 96 17% 2 213 4347 224 20.4 18 8% 16 644 9826 505 15.3 44 9% 12 723 7046 362 8.9 58 16% 5 258 3535 182 12.2 25 14% 1 525 4272 220 7.4 42 19% 4 308 4119 212 12.1 17 8% 9 252 3504 180 12.3 25 14% 6 227 3375 174 13.1 24 14% 7 257 3530 182 12.2 25 14% 10 194 2105 108 9.8 12 11% 8 405 3203 165 9.6 29 18% 14 659 3419 176 5.2 16 9% 13 300 8617 443 27.4 8 2% 17 285 4741 244 15.7 5 2% 21 582 10359 533 16.8 60 11% 22 640 10421 536 15.7 58 11% 23 962 17246 887 17.8 130 15% 24 546 10352 533 22.7 76 14% 25 831 26756 1376 34.2 110 8% 19 1315 23699 1219 15.1 104 9% 20 515 11282 580 17.4 88 15% 18 641 15710 808 23.3 34 4% 1-14 525 18364 945 32.8 47 5% 1-13 308 9443 486 28.2 28 6% 1-11 308 9443 486 28.2 28 6% 1-8 227 6937 357 27.3 19 5% 1-6 227 6937 357 27.3 19 5% 1-5 227 6937 357 27.3 19 5% 1-4 255 7621 392 26.9 20 5% 1-3 384 10473 539 27.5 43 8% 1-1 306 7625 392 23.8 23 6% Totals & Averages 14,943 286,382 14,732 1,351

Table C-2 MV Design Output SADB 57 F Max OA % 19% % OA CFM 2782 CFM Avg RADB 76.2 F

Zone

Table C-3 MV Occupied Cooling Zone Level Q V from V with OADB Q Peak Block QL Block AHU min V/Area F BTU/hr BTU/hr BTU/hr CFM CFM CFM/ft^2 97 286382 231969 22400 11933 11933 0.798 92 189393 153409 22400 7891 7891 0.528 87 125252 101454 22400 5219 5219 0.349 82 82833 67095 22400 3451 3451 0.231 77 54780 44372 22400 2283 2983 0.199 72 36228 29344 22400 1509 2983 0.199 67 23959 19406 22400 998 2983 0.199 62 15845 12834 22400 660 2983 0.199 57 10479 8488 22400 437 2983 0.199

Table C-4 MV Unoccupied Cooling Zone Level Q Q V V w/ OADB Q Peak Block Latent Block Min V/A F BTU/hr BTU/hr BTU/hr CFM CFM CFM/ft2 97 124561 100894 0 8493 8493 0.568 92 111251 90114 0 7585 7585 0.507 87 99364 80485 0 6775 6775 0.453 82 88747 71885 0 6051 6051 0.404 77 79265 64204 0 5404 5404 0.361 72 70795 57344 0 4827 4827 0.323 67 63231 51217 0 4311 4311 0.288 62 56475 45744 0 3851 3851 0.257 57 50440 40857 0 3439 3439 0.23

System

Table C-5 MV Occupied System Level V w/ OADB OAW OAW OARH V min XOA RADB QRH QPCC F gr/lb lb/lb % CFM CFM % F BTU/hr BTU/hr 97 126.9 0.0181 48 11933 11933 23% 76.2 0 120183 92 128.5 0.0184 56 7891 7891 35% 76.2 0 105159 87 108.6 0.0155 56 5219 5219 53% 76.2 0 90137 82 104.9 0.0150 64 3451 3451 81% 76.2 0 75114 77 88.4 0.0126 63 2283 2983 93% 71.7 14543 60091 72 67.1 0.0096 57 1509 2983 93% 66.7 30588 45068 67 59.8 0.0086 61 998 2983 93% 63.4 41199 30045 62 54.0 0.0077 65 660 2983 93% 61.3 48217 15022 57 43.4 0.0062 63 437 2983 93% 59.8 52858 0

Table C-5 Continued

QPCC OADB Latent MADB CCLADB CCEADB PHLADB QPHC QCS RAW F BTU/hr F F F F BTU/hr BTU/hr Lb/lb 97 170680 71.7 57 71.7 71.7 0 189936 0.00584 92 173716 69.4 57 69.4 69.4 0 106054 0.00604 87 135521 65.9 57 65.9 65.9 0 50580.3 0.00634 82 128376 60.7 57 60.7 60.7 0 13893.8 0.00680 77 96661.9 58.2 57 58.2 58.2 0 4174.94 0.00701 72 55621.4 58.2 57 58.2 58.2 0 4174.94 0.00701 67 41673.8 58.2 57 58.2 58.2 0 4174.94 0.00701 62 30436.1 58.2 57 58.2 58.2 0 4174.94 0.00701 57 10083.4 58.2 57 58.2 58.2 0 4174.94 0.00701

Table C-5 Continued OADB RAW MAW MAW MARH CCLAW QRH QCL QCT F gr/lb lb/lb gr/lb % lb/lb BTU/hr BTU/hr BTU/hr 97 40.9 0.00575 40.28 35 0.00546 0 187858 497976 92 42.3 0.00584 40.85 38 0.00546 0 188219 399433 87 44.4 0.00587 41.09 43 0.00546 0 145981 286698 82 47.5 0.00572 40.02 50 0.00546 0 132720 221728 77 49.0 0.00556 38.93 53 0.00546 14543 98172 162439 72 49.0 0.00556 38.93 53 0.00546 30588 57131 106375 67 49.0 0.00556 38.93 53 0.00546 41199 43184 77404 62 49.0 0.00556 38.93 53 0.00546 48217 31946 51144 57 49.0 0.00556 38.93 53 0.00546 52858 11593 15768

Table C-6 MV Unoccupied Cooling System Level V w/ QRH OADB MCWB OAW OAW OARH V min VOA RADB AHU F F gr/lb lb/lb % CFM CFM CFM F BTU/hr 97 79.8 126.9 0.0181 47.7 8493 8493 0% 66 18344 92 78.8 128.5 0.0184 56.3 7585 7585 0% 66 16384 87 74.4 108.6 0.0155 56.0 6775 6775 0% 66 14634 82 72.4 104.9 0.015 63.6 6051 6051 0% 66 13070 77 67.9 88.4 0.0126 63.3 5404 5404 0% 66 11674 72 61.9 67.1 0.0096 57.1 4827 4827 0% 66 10426 67 58.5 59.8 0.0086 60.5 4311 4311 0% 66 9312 62 55.1 54.0 0.0077 65.1 3851 3851 0% 66 8317 57 50.2 43.4 0.0062 62.8 3439 3439 0% 66 7428

Table C-6 Continued

Real OADB RADB MADB CCLADB CCEADB PHLADB QPHC QCS RARH F F F F F F BTU/hr BTU/hr % 97 68 68 55 68 68 0 119238 37 92 68 68 55 68 68 0 106498 37 87 68 68 55 68 68 0 95119 37 82 68 68 55 68 68 0 84955 37 77 68 68 55 68 68 0 75878 37 72 68 68 55 68 68 0 67770 37 67 68 68 55 68 68 0 60529 37 62 68 68 55 68 68 0 54062 37 57 68 68 55 68 68 0 48285 37

Table C-6 Continued OADB RAW RAW MAW MAW MARH CCLAW QRHC QCL QCT F lb/lb gr/lb lb/lb gr/lb % lb/lb BTU/hr BTU/hr BTU/hr 97 0.00504 35.3 0.00504 35.3 37 0.005044 18344 0 119238 92 0.00504 35.3 0.00504 35.3 37 0.005044 16384 0 106498 87 0.00504 35.3 0.00504 35.3 37 0.005044 14634 0 95119 82 0.00504 35.3 0.00504 35.3 37 0.005044 13070 0 84955 77 0.00504 35.3 0.00504 35.3 37 0.005044 11674 0 75878 72 0.00504 35.3 0.00504 35.3 37 0.005044 10426 0 67770 67 0.00504 35.3 0.00504 35.3 37 0.005044 9312 0 60529 62 0.00504 35.3 0.00504 35.3 37 0.005044 8317 0 54062 57 0.00504 35.3 0.00504 35.3 37 0.005044 7428 0 48285

Plant

Figure C-1 Part Load Operation Data for 50 ton Trane Chiller

Figure C-2 Effect of Ambient Temperature on 50 ton Trane Chiller

Table C-7 MV Occupied Cooling Plant Level % Full Real OADB Freq QCT QCT Cap EER Factor EER Factor EER F Hours BTU/hr Tons Load Temp 97 0 497976 41.5 59.90% 10 1 9.7 1 9.7 92 94.1 399433 33.3 48.00% 10.4 1 10.4 1 10.8 87 363.6 286698 23.9 34.50% 10.1 1 11.1 1.1 11.2 82 480 221728 18.5 26.70% 9.5 0.9 11.4 1.1 10.8 77 292.9 162439 13.5 19.50% 8.7 0.9 11.4 1.1 10 72 372.2 106375 8.9 12.80% 7.8 0.8 11.4 1.1 8.9 67 400.6 77405 6.4 9.30% 7.2 0.7 11.4 1.1 8.2 62 299.5 51144 4.3 6.20% 6.7 0.7 11.4 1.1 7.6 57 156.3 15769 1.3 1.90% 5.8 0.6 11.4 1.1 6.7

Table C-7 Continued OADB Chiller Fan Reheat Tot SP in F kW kWh H2O kW kWh BTU/hr kW kWh kW kWh 97 51.3 0 4.00 10.5 0 0 0 0 61.8 0 92 36.9 3476 1.75 3.0 286 0 0 0 39.9 3762 87 25.6 9317 0.77 0.9 319 0 0 0 26.5 9636 82 20.4 9832 0.33 0.3 122 0 0 0 20.7 9954 77 16.3 4781 0.25 0.2 48 14543 4.2 1249 20.7 6077 72 11.9 4455 0.25 0.2 61 30588 8.9 3337 21.1 7853 67 9.3 3760 0.25 0.2 66 41199 12.0 4837 21.6 8664 62 6.7 2013 0.25 0.2 49 48217 14.1 4232 21.0 6293 57 2.3 369 0.25 0.2 26 52858 15.4 2420 18.0 2815

Table C-8 MV Unoccupied Cooling Plant Level % Full Real OADB Frequency QCT QCT Cap EER Factor EER Factor EER F Hours BTU/hr Tons Load Temp 97 0 119238 9.9 14.3% 8.0 0.8 9.7 1.0 7.8 92 60.9 106498 8.8 12.8% 7.8 0.8 10.4 1.0 8.1 87 397.4 95119 7.9 11.4% 7.6 0.8 11.1 1.1 8.4 82 558.0 84955 7.0 10.2% 7.4 0.7 11.9 1.2 8.7 77 566.1 75878 6.3 9.1% 7.2 0.7 12.6 1.3 9.0 72 1022.8 67770 5.6 8.1% 7.0 0.7 13.3 1.3 9.3 67 658.4 60529 5.0 7.3% 6.9 0.7 14.0 1.4 9.6 62 766.5 54062 4.5 6.5% 6.7 0.7 14.8 1.5 9.9 57 454.8 48285 4.0 5.8% 6.6 0.7 15.5 1.5 10.2

Table C-8 Continued OADB Chiller Fan Reheat Tot F kW kWh SP in H2O kW kWh BTU/hr kW kWh kW kWh 97 15.3 0 2.03 5.3 0 18344 5.3 0 26 0 92 13.1 800 1.62 2.8 171 16384 4.8 292 20.7 1264 87 11.2 4486 1.29 1.5 588 14634 4.2 1704 17 6779 82 9.7 5424 1.03 0.78 436 13070 3.8 2137 14.3 7997 77 8.3 4749 0.82 0.54 305 11674 3.4 1937 12.3 6990 72 7.2 7420 0.65 0.43 439 10426 3 3125 10.7 10985 67 6.2 4136 0.52 0.34 226 9312 2.7 1797 9.3 6159 62 5.4 4175 0.42 0.27 210 8317 2.4 1868 8.1 6253 57 4.7 2149 0.33 0.22 99 7428 2.1 990 7.1 3239

Heating

Table C-9 MV Heating Design and Zone Level Zone Name Area QOE QL Bldg U Q Loss VOA Q need QBB QBB*1.1 ft2 BTU/hr BTU/hr BTU/hr*F BTU/hr CFM BTU/hr BTU/hr BTU/hr 3 894 6498 4638 0 0 108 1519 0 0 2 213 3242 1105 0 0 42 593 0 0 16 644 6485 3341 0 0 95 1340 0 0 12 723 3295 3751 0 0 68 961 0 0 5 258 2197 1338 0 0 34 482 0 0 1 525 1548 2724 0 0 42 583 0 0 4 308 2521 1598 0 0 40 562 0 0 9 252 2197 1307 0 0 34 478 0 0 6 227 2197 1178 0 0 33 460 0 0 7 257 2197 1333 0 0 34 481 0 0 10 194 1098 1006 0 0 20 287 0 0 8 405 2197 1006 0 0 31 437 0 0 14 659 0 3419 0 0 33 466 0 0 13 300 7060 1556 0 0 84 1175 0 0 17 285 900 1401 54 1959 46 647 305 336 21 582 5322 2860 49 1781 101 1413 300 330 22 640 4257 3145 68 2478 101 1421 300 330 23 962 2579 4728 57 2058 168 2352 300 330 24 546 720 2385 54 1978 101 1412 300 330 25 831 2197 4311 170 6134 260 3649 3276 3603 19 1315 10800 6463 113 4077 230 3232 300 330 20 515 6464 2531 51 1870 110 1539 300 330 18 641 2988 2800 120 4333 153 2143 688 756 1-14 525 2988 2259 170 6155 178 2505 3413 3754 1-13 308 1992 1346 62 2256 92 1288 300 330 1-11 308 1992 1346 62 2256 92 1288 300 330 1-08 227 1446 992 46 1662 67 946 300 330 1-06 227 1446 992 46 1662 67 946 300 330 1-05 227 1446 992 46 1662 67 946 300 330 1-04 255 1446 1114 51 1870 74 1039 350 385 1-03 384 2346 1678 118 4264 102 1428 1669 1836 1-01 306 1514 1474 79 2850 74 1040 901 991 Totals & Averages 14,943 95,579 72,116 1,425 51,308 2782 13,901 15,291

Table C-10 MV Heating System Level OADB QLOSS QINT QBB V RADB MADB QRH QBB F BTU/hr BTU/hr BTU/hr CFM F F BTU/hr kW 52 22804 167,695 1181 2782 68 52 9014 0.346 47 29930 167,695 5760 2782 68 47 24037 1.68 42 37056 167,695 10339 2782 68 42 39059 3.03 37 44182 167,695 14918 2782 68 37 54082 4.37 32 51308 167,695 19497 2782 68 32 69105 5.71

Table C-11 MV Heating Plant Level OADB Frequency Fan Reheat/BB Tot SP in F Hours H2O kW kWh BTU/hr kW kWh kW kWh 52 200.9 0.22 0.14 29 9014 2.9 600 3.1 629 47 139.6 0.22 0.14 20 24037 8.7 1219 8.8 1239 42 54.4 0.22 0.14 8 39059 14.4 787 14.6 795 37 13.8 0.22 0.14 2 54082 20.2 280 20.3 282

APPENDIX D WEATHER DATA

Table D-1 Monthly Design Day Weather Data Jan 31 Feb 28 Mar 31 Bin Freq Monthly Bin Freq Monthly Bin Freq Monthly 0 0 0 0 0 0 0 0 0 5 0 0 5 0 0 5 0 0 10 0 0 10 0 0 10 0 0 15 0 0 15 0 0 15 0 0 20 0 0 20 0 0 20 0 0 25 0 0 25 0 0 25 0 0 30 0 0 30 0 0 30 0 0 35 0 0 35 0 0 35 0 0 40 0 0 40 0 0 40 0 0 45 4 124 45 3 84 45 0 0 50 3 93 50 5 140 50 5 155 55 5 155 55 3 84 55 5 155 60 2 62 60 2 56 60 3 93 65 4 124 65 3 84 65 3 93 70 6 186 70 4 112 70 4 124 75 0 0 75 4 112 75 4 124 80 0 0 80 0 0 80 0 0 85 0 0 85 0 0 85 0 0 90 0 0 90 0 0 90 0 0 95 0 0 95 0 0 95 0 0 100 0 0 100 0 0 100 0 0 More 0 0 More 0 0 More 0 0 24 24 24

Table D-1 Continued Apr 30 May 31 Jun 30 Bin Freq Monthly Bin Freq Monthly Bin Freq Monthly 0 0 0 0 0 0 0 0 0 5 0 0 5 0 0 5 0 0 10 0 0 10 0 0 10 0 0 15 0 0 15 0 0 15 0 0 20 0 0 20 0 0 20 0 0 25 0 0 25 0 0 25 0 0 30 0 0 30 0 0 30 0 0 35 0 0 35 0 0 35 0 0 40 0 0 40 0 0 40 0 0 45 0 0 45 0 0 45 0 0 50 0 0 50 0 0 50 0 0 55 0 0 55 0 0 55 0 0 60 1 30 60 1 31 60 0 0 65 6 180 65 6 186 65 0 0 70 4 120 70 4 124 70 4 120 75 4 120 75 3 93 75 5 150 80 3 90 80 3 93 80 3 90 85 6 180 85 7 217 85 5 150 90 0 0 90 0 0 90 7 210 95 0 0 95 0 0 95 0 0 100 0 0 100 0 0 100 0 0 More 0 0 More 0 0 More 0 0 24 24 24

Table D-1 Continued July 31 Aug 31 Sept 30 Bin Freq Monthly Bin Freq Monthly Bin Freq Monthly 0 0 0 0 0 0 0 0 0 5 0 0 5 0 0 5 0 0 10 0 0 10 0 0 10 0 0 15 0 0 15 0 0 15 0 0 20 0 0 20 0 0 20 0 0 25 0 0 25 0 0 25 0 0 30 0 0 30 0 0 30 0 0 35 0 0 35 0 0 35 0 0 40 0 0 40 0 0 40 0 0 45 0 0 45 0 0 45 0 0 50 0 0 50 0 0 50 0 0 55 0 0 55 0 0 55 0 0 60 0 0 60 0 0 60 0 0 65 0 0 65 0 0 65 0 0 70 0 0 70 0 0 70 0 0 75 7 217 75 6 186 75 9 270 80 5 155 80 5 155 80 3 90 85 5 155 85 4 124 85 5 150 90 7 217 90 4 124 90 7 210 95 0 0 95 5 155 95 0 0 100 0 0 100 0 0 100 0 0 More 0 0 More 0 0 More 0 0 24 24 24

Table D-1 Continued Oct 31 Nov 30 Dec 31 Bin Freq Monthly Bin Freq Monthly Bin Freq Monthly 0 0 0 0 0 0 0 0 0 5 0 0 5 0 0 5 0 0 10 0 0 10 0 0 10 0 0 15 0 0 15 0 0 15 0 0 20 0 0 20 0 0 20 0 0 25 0 0 25 0 0 25 0 0 30 0 0 30 0 0 30 0 0 35 0 0 35 0 0 35 0 0 40 0 0 40 0 0 40 5 155 45 0 0 45 2 60 45 4 124 50 0 0 50 6 180 50 3 93 55 1 31 55 3 90 55 3 93 60 6 186 60 2 60 60 3 93 65 3 93 65 4 120 65 6 186 70 3 93 70 6 180 70 0 0 75 3 93 75 1 30 75 0 0 80 6 186 80 0 0 80 0 0 85 2 62 85 0 0 85 0 0 90 0 0 90 0 0 90 0 0 95 0 0 95 0 0 95 0 0 100 0 0 100 0 0 100 0 0 More 0 0 More 0 0 More 0 0 24 24 24

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BIOGRAPHICAL SKETCH

Lane W. Burt was born in Charlotte, NC. He began working for his father in the HVAC industry during summers away from school at age 14, learning in a hands-on manner about the construction and mechanical contracting businesses. He graduated from North Carolina State

University in 2005 and opted to pursue a graduate education at the University of Florida. At

Florida he worked under Dr. H.A. (Skip) Ingley in the Alternative Energy Laboratory focusing on building energy consumption and “green” or high-performance buildings.