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

Transcription

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

2 2007 Lane W. Burt 2

3 To my Dad, whose advice was invaluable 3

4 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. 4

5 TABLE OF CONTENTS ACKNOWLEDGMENTS...4 LIST OF TABLES...7 LIST OF FIGURES...9 ABSTRACT...10 CHAPTER 1 INTRODUCTION AND BACKGROUND...11 page Motivation...11 Displacement Ventilation...12 Application 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 BUILDING FOR STUDY...25 Purpose...25 The Building...25 Specifications...25 Zoning DISPLACEMENT VENTILATION DESIGN...29 System Design...29 Design Specifications...29 Displacement Ventilation...29 Design Process...29 Application...33 Mixing Ventilation ENERGY ANALYSIS

6 Building Load Profile...38 Bin Weather Data...39 All Zone Level...39 System Level...40 Plant Level...40 Analysis COST ANALYSIS...48 Life Cycle Cost Analysis...48 Electricity Cost...48 First Costs 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

7 LIST OF TABLES Table page 3-1 Room Zoning Zone Loads DV Design Bin Weather Data DV Occupied Zone Level MV Occupied All Zone Level DV Occupied System Level DV Occupied Plant Level Energy Consumption Checksums DV and MV Cooling Energy Consumption Comparison DV and MV Energy Consumption Comparison Monthly Electric Demand Electric Rates and Costs First Cost Unit Inputs System First Costs DV Diffuser Data 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

8 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

9 LIST OF FIGURES Figure page 1-1 Mixing vs. Displacement Ventilation First Floor Plan 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

10 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 Chair: H. A. Ingley Major: Mechanical Engineering By Lane W. Burt December 2007 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. 10

11 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 CO 2 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 11

12 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. 12

13 The bulk of the research on displacement ventilation has been conducted in Scandinavian countries, where it has been used commonly since its introduction in 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 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 13

14 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. 14

15 Figure 1-1 Mixing vs. Displacement Ventilation 15

16 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 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 16

17 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*ft 2 (40 17

18 W/m 2 ) 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*ft 2 (120 W/m 2 ) 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 18

19 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, 19

20 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 20

21 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. 21

22 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 22

23 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. 23

24 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

25 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*ft 2 * F (0.62 W/m 2 *C) based on the original load calculation from the design engineer. The U factor of the floor was 0.21 BTU/hr*ft 2 * F (1.21 W/m 2 *C). The glass possesses very high performance attributes with a U factor and shading coefficient of

26 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. 26

27 Figure 3-1 First Floor Plan 27

28 Table 3-1 Room Zoning Zone Name Room Load Name Rm # Zone Name Room Load Name Rm # 3 Corridor Workroom Board Room R Support 102A 2 Toilet Receiving Coffee Corridor 4 ext Corridor 4 int 154, Marketing West Coffee Bar Marketing East Corridor Research Lib West Comp Dept Office Research Lib Mid Comp Dept Cubicles Research Lib East Accounts Payable Lounge Consultant Meeting Room Corridor 1s Unassigned Office 157A 1 Exec Asst 107, 112, 108, Unassigned Office Exec Asst 121, Storage Storage Research Office Human Resources Marketing Human Resources Marketing Interview Dir Marketing Recruiter Dir Edit Services Recruiter Dir Hmn Resources Human Resources Dir Inside Sales Unassigned Office Dir Outside Sales Storage VP Sales Corridor 1n CFO Unassigned Office COO Men's 164, President Custodial Mtg Rm & Dining Women's 167, Chairman

29 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*ft 2 (120 W/m 2 ) 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. 29

30 Calculate the air changes per hour: Based on heat gains using equation 4-1, (0.295QOE QL QEX ) n = ΔTρC HA p (4-1) With ΔT = delta temperature between foot and head level (taken to be 3.6F or 2C) ρ = density of air (lb/ft 3 ) C p = specific heat of air (BTU/lb*F) H = height of ceiling (ft) A = floor area (ft 2 ) 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& C = (4-2) 60 Calculate the flow rate of fresh air: This value is calculated by using the ASHRAE required flow rate and the ventilation effectiveness, η. The actual effectiveness can be calculated as follows in equation 4-3, 0.28n ( QOE + 0.4QL + 0.5QEX ) η = 3.4(1 e ) (4-3) ( QOE + QL + QEX ) 30

31 The ventilation effectiveness is a measure of how much fresh air is needed to attain acceptable indoor air quality. ASHRAE Std states that typical ventilation effectiveness for a DV system is 1.2 while mixing is assumed to be , The outdoor air ventilation rate for the DV system can then be calculated from equation V 62.1 VOA = (4-4) η With the variable V62.1 representing the ventilation rate needed for the occupied space based on ASHRAE Std 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 = max( VOA, VC) (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 (60VρC p ) 1 1 θ f = *( + ) + 1 (4-6) A α R α CF With α R = radiative heat transfer coefficient α CF = convective heat transfer coefficient Both coefficients have units of BTU/(hr*ft 2 * F) and are approximately 1 as stated by ASHRAE. The nrequired supply temperature based on dimensionless temperature can be calculated from equation 4-7, 31

32 ( θ f ( QOE + QL + QEX )) SADB = ( TSET ΔThf ) (4-7) 60ρC V With TSET = design room temperature (i.e. 75 F, 23.9 C) ΔT hf = 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 p 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. 4-8 The return air temperature (RADB) can be calculated from the energy balance in equation Q RADB = SADB + (4-8) 60ρC V p 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, A = V / 40 fpm (4-10) 32

33 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*ft 2 (120 W/m 2 ) 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. 33

34 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. 34

35 Table 4-1 Zone Loads QL (lights) QEX (exterior) QOE (internal) Q Q/A Area (ft 2 ) (BTU/hr) (BTU/hr) (BTU/hr) (BTU/hr) (BTU/ft 2* hr) Zone , , , , , , , , , , , , , , , , , , , , , , , , , , , ,401 2, DV Marketing West 582 2,860 2,177 5, DV Marketing East 640 3,145 3,019 4, DV Research West 962 4,728 9,939 2, DV Research Mid 546 2,385 7, DV Research East 831 4,311 20,248 2, ,463 6,436 10, ,531 2,286 6, ,800 9,922 2, ,259 13,118 2, ,346 6,106 1, ,346 6,106 1, ,499 1, ,499 1, ,499 1, ,114 5,061 1, ,678 6,449 2, ,474 4,636 1, Total 14,943 72, ,687 95,

36 Table 4-2 DV Design Cooling Flowrate Effectiveness Outdoor Air Outdoor Air DV Flowrate Temp Gradient Supply Temp Zone n Vc η V62.1 VOA V θf SADB Name (ACPH) (cfm) (cfm) (cfm) (cfm) (F) Totals & Averages

37 Table 4-2 Continued V recalc Return Temp OA Fraction (cfm) RADB XOA Diffuser Zone Name (F) Area (ft 2) % % % % % % % % % % % % % % % % % % % % % % % % % % % % % % % % 10.4 Totals & Averages

38 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. y x = e (5-1) y x = e (5-2) y = x (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 38

39 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. 39

40 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. 40

41 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. 41

42 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. 42

43 Figure 5-1 Building Load Profile Table 5-1 Bin Weather Data Bin OADB (F) Occupied (hrs) Unoccupied (hrs) More

44 Table 5-2 DV Occupied Zone Level Outdoor Air Temp Peak Load Block Load Latent Load Peak Airflow Block Airflow Real airflow w/ AHU min OADB QP QB QLA V V V CFM per square foot F BTU/hr BTU/hr BTU/hr CFM CFM CFM CFM/FT Table 5-3 MV Occupied All Zone Level Outdoor Air Temp Peak Load Block Load Latent Load Peak Airflow Block Airflow Real airflow w/ AHU min OADB QP QB QLA V V V CFM per square foot F BTU/hr BTU/hr BTU/hr CFM CFM CFM CFM/FT

45 Table 5-4 DV Occupied System Level Outdoor Outdoor Air Temp Mean Coincident Wet Bulb Air Humidity Ration Airflow Airflow w/ minimum Outdoor Air Fraction OADB MCWB OAW V V XOA Return Dry Bulb RADB Pre- Cooling Coil Load QPCC Mixing Air Dry Bulb MADB F F lb/lb CFM CFM % F BTU/hr F % % % % % % % % % Table 5-4 Continued Outdoor Air Temp Sensible Coil Load Return Air Humidity Ratio Mixing Air Humidity Ratio Cooling Coil Leaving Humidity Reheat Coil Load Coil Latent Load Total Coil Load OADB QCS RAW MAW CCLAW QRHC QCL QCT F BTU/hr lb/lb lb/lb lb/lb BTU/hr BTU/hr BTU/hr

46 Table 5-5 DV Occupied Plant Level Outdoor Air Temp Total Coil Load Total Coil Load Chiller Fan Reheat OADB Freq QCT QCT % Load EER Power Power Power Total Power Total F Hours BTU/hr Tons kw kw kw kw kwh % % % % % % % % % % % % % Table 5-6 Energy Consumption Mode Operation kwh Occupied Unocc DV Total Occ Unocc MV Total Table 5-7 Checksums Checksums DV MV Energy kwh/ft^ CFM/ft^ Cooling ft^2/ton Heating BTU/hr*ft^

47 Table 5-8 DV and MV Cooling Energy Consumption Comparison DV kwh MV kwh Chiller Chiller Fan 1114 Fan 977 Reheat Reheat Tot Tot Table 5-9 DV and MV Energy Consumption Comparison Mode DV MV % change kwh kwh Occupied Cooling % Occupied Heating % Unoccupied Cooling % All Cooling % Total % 47

48 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, 48

49 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 49

50 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. 50

51 Table 6-1 Monthly Electric Demand Month Demand (kw) $ DV MV DV MV Jan $170 $220 Feb $202 $202 Mar $202 $202 Apr $192 $233 May $192 $233 Jun $308 $353 Jul $308 $353 Aug $503 $551 Sep $308 $353 Oct $192 $233 Nov $202 $202 Dec $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 $0.06 per kwh for first 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 CFM Tons supply grills VAV boxes heaters RA grills 51

52 Table 6-4 System First Costs Category Item Rate Units DV MV Units Duct Standard 0.6 lbs/cfm 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 Grills from supplier $20,281 $ - $ fro DV grills 400 $/grill $18,800 $ - $ Chases 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 Heaters 1 kw 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 Face Areas Max 40 fpm # needed Drop Size Cost ft2 CFM in $ $6, $8, x6 $2, x8 $1, x8 $1, $20,281 52

53 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, $ - $11,520 $ - $11, $ - $11,184 $ - $11, $ - $10,858 $ - $10, $ - $10,542 $ - $10, $ - $10,235 $ - $10, $ - $9,937 $ - $9, $ - $9,647 $ - $9, $ - $9,366 $ - $9, $ - $9,094 $ - $9, $ - $8,829 $ - $8, $ - $8,572 $ - $8,514 Total (present worth) $417,342 $382,715 53

54 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 54

55 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. 55

56 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. 56

57 APPENDIX A BUILDING FLOORPLAN Figure A-1 Building Floorplan 57

58 APPENDIX B DISPLACEMENT VENTILATION SPREADSHEETS Design Table B-1 Input Parameters for DV Design aoe Thf (F) 3.6 H (ft) 12 TSET, T-head (F) 75 al ρ (lb/ft3) αr 1 T-foot (F) 71.4 aex 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 58

59 Table B-3 DV Design Sizing Zone Name # Area Rooms A QOE QL QEX Q Total Load Density Q/A ft 2 BTU/hr BTU/hr BTU/hr BTU/hr BTU/hr*ft 2 W/m Totals & Averages 14,943 73,532 72, , ,335 Q/A 59

60 Table B-3 Continued Cooling Flow Rate VC V Effective ness η Zone Name n n Min VC Min OA Flow Rate V62 XOA VOA ACPH ACPH CFM CFM CFM CFM % % % % % % % % % % % % % % % % % % % % % % % % % % % % % % % % 15 Totals & Averages 13,693 8,038 1,

61 Table B-3 Continued Zone Name V θf SADBi RADB SADB- SADBi VOA Min V Min Min % CFM F F F CFM CFM % % % % % % % % % % % % % % % % % % % % % % % % % % % % % % % % Totals & Averages 13,

62 Table B-3 Continued Zone Name V recalc Thf recalc Tf recalc Th recalc θf recalc RADB Recalc XOA RADB*V Diffuser Area A CFM F F C F % CFM*F ft % % % % % % % % % % % % % % % % % % % % % % % % % % % % % % % % Totals & Averages %

63 Zone Table B-4 DV Occupied Zone Inputs Altitude 155 ft Peak V CFM Block Load BTU/hr Area 14,943 ft^2 Min V 2887 CFM Block V CFM Q 264,335 BTU/hr SADB 68.7 F AHU Size CFM Eqn Coeff a TSET 75 F AHU Min 25% % Y=ae^bx b # People 112 AHU Min CFM Table B-5 DV Occupied All Zones Q OADB Q Peak Zones Block Zones Q Zone Latent Peak V Block V V w/ min V/A F BTU/hr BTU/hr BTU/hr CFM CFM CFM CFM/ft Table B-6 DV Unoccupied Zone Inputs Altitude 155 ft Peak V 6664 CFM Block Load 124,561 BTU/hr Area ft^2 Min V 0 CFM Block V 6664 CFM Q 124,561 BTU/hr SADB 59 F AHU Size CFM Eqn Coeff a TSET 68 F AHU Min 0.25 % Y=ae^bx b # People 0 AHU Min 3402 CFM 63

64 Table B-7 DV Unoccupied All Zones OADB Q peak zones Q block zones Q Zone Latent Peak V Block V w/ min V/A F BTU/hr BTU/hr BTU/hr CFM CFM CFM/ft System Table B-8 DV Occupied Cooling System Inputs Altitude 155 ft V Peak 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 lb/lb TSET 75 Table B-9 DV Occupied Cooling System Level Outdoor Air Dry Bulb Mean Coincident Wet Bulb Outdoor Air Humidity Ratio Outdoor Air Humidity Ratio Outdoor Air Relative Humidity Return Air Dry Bulb (without Real Total w. min % OA reheat) OADB MCWB OAW OAW OARH V V XOA RADB QRH from AHU min F F gr/lb lb/lb % CFM CFM % F BTU/hr % % % % % % % % %

65 Table B-9 Continued Outdoor Air Dry Bulb Real RADB Pre Cooling Sensible Pre Cooling Latent Mix Air Dry Bulb Cooling Coil Leaving Dry Bulb Cooling Coil Entering Air DB Preheat Leaving Dry Bulb Coil Load Return Air Relative Humidity OADB RADB QPCC QPCC MADB CCLADB CCEADB PHLADB QCS RARH F F BTU/hr BTU/hr F F F F BTU/hr % Table B-9 Continued Outdoor Air Dry Bulb Return Air Humidity Ratio Return Air Humidity Ratio Mixing Air Humidity Ratio Mixing Air Humidity Ratio Mixing Air Relative Humidity Cooling Coil Leaving Humidity Ratio Re Heat Latent Load OADB RAW RAW MAW MAW MARH CCLAW QRHC QCL QCT Total Load F lb/lb gr/lb lb/lb gr/lb % lb/lb BTU/hr BTU/hr BTU/hr

66 Table B-10 DV Unoccupied Cooling System Level Outdoor Air Dry Bulb Mean Coincident Wet Bulb Outdoor Air Humidity Ratio Outdoor Air Humidity Ratio Outdoor Air Relative Humidity Total Real w. min Outdoor Air Return Air Dry Bulb (without reheat) OADB MCWB OAW OAW OARH V V XOA RADB QRH from AHU min F F gr/lb lb/lb % CFM CFM % F BTU/hr % % % % % % % % % Table B-10 Continued Outdoor Air Dry Bulb Real RADB Pre Cooling Sensible Mix Air Dry Bulb Cooling Coil Leaving Dry Bulb Cooling Coil Entering Air DB Preheat Leaving Dry Bulb Pre Heat Coil Load Return Air Relative Humidity OADB RADB QPCC MADB CCLADB CCEADB PHLADB QPHC QCS RARH F F BTU/hr F F F F BTU/hr BTU/hr %

67 Table B-10 Continued Outdoor Air Dry Bulb Return Air Humidity Ratio Return Air Humidity Ratio Mixing Air Humidity Ratio Mixing Air Humidity Ratio Mixing Air Relative Humidity Cooling Coil Leaving Humidity Ratio Re Heat Latent Load OADB RAW RAW MAW MAW MARH CCLAW QRHC QCL QCT Total Load F lb/lb gr/lb lb/lb gr/lb % lb/lb BTU/hr BTU/hr BTU/hr Plant Figure B-1 Part Load Operation Data for 40 ton Trane Chiller 67