Reference

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cations AlllOt Ml c)2 lf c

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NBSIR 82-2489

Strategies for Energy

Conservation in Small Office

Buildings

U.S. DEPARTMENT OF COMMERCE

National Bureau of Standards

National Engineering Laboratory

Center for Building Technology

Building Equipment Division

Washington, DC 20234

June 1 982

— QC

100

.U96

82-2489

1982

Prepared for Office of Buildings Energy Research

__and Development

ffice of Conservation and Renewable Energy

.S. Department of Energy

a, or btandaadb

uaxAMr

JUL ^ i

CL NBSIR 82-2489

STRATEGIES FOR ENERGY

CONSERVATION IN SMALL OFFICE

BUILDINGS

Walter H. Parken

James Y. Kao

George E. Kelly

U.S. DEPARTMENT OF COMMERCE

National Bureau of Standards

National Engineering Laboratory

Center for Building Technology

Building Equipment Division

Washington, DC 20234

June 1982

Prepared for Office of Buildings Energy Research

and Development

Office of Conservation and Renewable Energy

U.S. Department of Energy

U.S. DEPARTMENT OF COMMERCE, Malcolm Baldrige, Secretary

NATIONAL BUREAU OF STANDARDS, Ernest Ambler, Director

ABSTRACT

A comparative analysis is made of the thermal performance of a small office

building using various HVAC systems and commonly emDloyed strategies. The com- parisons are made for seven geographical locations representing xyide climatic

variations within the continental United States.

Results were obtained for fan, space heating hot water, and chilled water energv

consumption through hour-by-hour simulations using the BT.AST computer program.

A small office building model was used in the simulations along with several

HVAC systems; a constant volume reheat unit (serving the entire building), dual

constant volume reheat units (serving separate zones of the building), and a variable air volume reheat unit. The strategies investigated included simply

air temperature reset (constant, zone-controlled, and outdoor air-controlled),

economy cycles (temperature and enthalpy), continuous conditioning versus con- ditioning only during occupied hours, changes in reheat set point temperature,

and changes in minimum variable air volume ratio.

For comparable control strategies, the variable-air volume terminal reheat

system exhibited the least energy consumption for chilled water, hot water and

circulating fan. The system incorporating two independently-operating constant

volume terminal reheat units ranked second in energy consumption while the

single constant volume terminal reheat unit ranked last. Changes in thermal

performance resulting from implementing one strategy in place of or in combination with another were found to vary significantly by climate and the type of

HVAC system employed.

Key words: Building control strategies; building energy conservation; building

thermal performance; HVAC.

iii
PREFACE

This report is one of a series documenting NBS research and analysis efforts in

developing energy and cost data to support the Department of Energy/National

Bureau of Standards Building Energy Conservation Criteria Program. The work

reported in this document was performed under the Energy Analysis of Control

Strategies project; a part of the Controls Program element managed by Brookhaven National Laboratory for DoE. The NBS effort was supported by DoE/NBS

Task Order A008-BCS under Interagency Agreement No. EA77A 01-5010.

iv
TABLE OF CONTENTS

Pare

ABSTRACT ill

PREFACE iv

LIST OF FIGURES vi

LIST OF TABLES vii

1. INTRODUCTION 1 2. PURPOSE AND SCOPE ...

3. STIMULATION PROCEDURE

3.1 THE BLAST PROGRAM 3 3.2 SMALL OFFICE BUILDING 3 3.3 WEATHER DATA 4

4. SIMULATION OF HVAC SYSTEMS AND STRATEGIES 5 4.1 SPACE CONDITIONING 5 4.2 CONSTANT VOLUME TERMINAL REHEAT SYSTEM SERVING THE ENTIRE

BUILDING 6

4.2.1 CV System Strategies 6

4.3 SYSTEM INCORPORATING TOO CONSTANT VOLUME TERMINAL REHEAT

UNITS 8

4.3.1 CV-CV System Strategies 9

4.4 VARIABLE AIR VOLUME TERMINAL REHEAT SYSTEM 11

4.4.1 VAV System Strategies 11

5. RESULTS AND DISCUSSIONS 14

5.1 COMPARISON OF SYSTEMS 14

5.2 COMPARISON OF STRATEGIES 16

6. SUMMARY 24

6.1

EVALUATION OF SYSTEMS AND STRATEGIES 27

APPENDIX A: RESULTS FROM SIMULATIONS OF ADDITIONAL STRATEGIES A-l

v

U)

NJ
LIST OF FIGURES

Figure Page

1. Small office building model 29

2. Internal heat gain schedules . 30

3. Annual heating and cooling energy consumption of three HVAC

systems utilizing the base case strategy ( A— 1 ) for a small

office building 31

4. Comparative annual heating and cooling energy consumption for a VAV system using the base case strategy with CV and CV-CV system

utilizing their representative best case strategies 32

5. Percent savings in chilled water and percent increases in hot

water energy consumption for three HVAC systems using enthalpy

economy cycle 33

6. Percent savings in chilled water and percent increases in hot

water energy consumption for three HVAC systems using temperature

economy cycle 34

7. Chilled water and hot water energy savings achieved for strategies

employing zone-controlled cold deck as compared to a fixed cold

deck temperature 35

vi
LIST OF TABLES

Table Page

1. Climatological Data for Seven Cities 16

2. Comparative Annual Energy Consumption of Three HVAC Systems and

Selected Strategies for a Small Office Building: Lake Charles .... 37

3. Comparative Annual Energy Consumption of Three HVAC Systems and

Selected Strategies for a Small Office Building: Nashville ....... 38

4. Comparative Annual Energy Consumption of Three HVAC Systems and

Selected Strategies for a Small Office Building: Washington DC ... 39

5. Comparative Annual Energy Consumption of Three HVAC Systems and

Selected Strategies for a Small Office Building: Dodge City 40

6. Comparative Annual Energy Consumption of Three HVAC Systems and

Selected Strategies for a Small Office Building: Madison 41

7. Comparative Annual Energy Consumption of Three HVAC Systems and

Selected Strategies for a Small Office Building: Seattle 42

8. Comparative Annual Energy Consumption of Three HVAC Systems and

Selected Strategies for a Small Office Building: Santa Maria ..... 43

A-l. Comparison of Energy Consumption for VAV Systems with Mininum Air

Volume Ratios of 0.2 and 0.4 for Seven Cities A-4

vii
„ . , . -
1. INTRODUCTION

Approximately 40 percent of the energy used annually in the United States is

consumed in buildings. Reduced building energy consumption can therefore have

a large impact in both public and private sectors and permit reallocation of scarce resources. The necessity for conserving energy has been recognized and

attempts to promote conservation have taken many forms, such as federallymandated emergency building temperature restrictions, proposed new standards

for energy conservation in new buildings, and development of more energyefficient heating and cooling equipment.

Recently, more energy efficient architectural designs have been employed in new commercial buildings. Examples include reduced glass area, utilizing

reflective and insulating glass, and reorienting glass surfaces to take advantage of natural daylighting and reducing summer solar transmission heat gain.

The energy savings due to efficient building shell design, however, can be

negated by high energy consumption due to improper selection, sizing, and

operation of HVAC equipment.

Computerized energy management and control systems (EMCS) are also gaining

popularity as an effective means to reduce building energy usage. Achieving

energy savings with these systems requires a thorough knowledge of HVAC equipment performance and the utilization of effective strategies for controlling

the building environment. The proper selection of HVAC systems and utilization

of effective strategies plays a key role in reducing heating and cooling costs

in new as well as existing buildings.

1
2. PURPOSE AND SCOPE

This is one of several reports intended to document the energy savings

potential of HVAC systems and strategies that can be employed in public

buildings. The subject of this report is the energy-savings potential of HVAC

systems and strategies employed in typical small office buildings.

The report is intended to be used as a general guide for designers, consultants,

and building owers or operators desiring to retrofit or install new HVAC systems

in small office buildings. This information can assist in determining from the

many possible combinations of systems and strategies those which are energyefficient and cost-effective. Once the choices of systems and strategies have

been narrowed, a more detailed analysis can be made incorporating the unique

aspects of a particular building and its intended use. In addition, it is

hoped that this report will create an awareness of the sensitivity of energy

cost to type of HVAC system installed and how it is operated.

Scope

In evaluating the thermal performance of selected HVAC systems, simulations were

made using the Building Loads Analysis System Thermodynamics (BLAST) * computer

program. This report focuses on three types of terminal reheat systems: a variable-air volume (VAV) system, two constant volume terminal reheat units

(CV-CV) serving separate zones of the building, and a single constant volume

(CV) system. These systems were selected since they are commonly employed in new and existing small office buildings and would be expected to show significant performance changes depending on climate and type of control strategy

employed. Induction units were not investigated due to the limitations in the

computer program selected for the simulations.

A small office building model was utilized to simulate the expected heating and

cooling loads imposed on the various selected systems and strategies. Since

the building loads as well as system performance is dependent on climate,

results we re obtained from weather data for seven cities.

Due to the impracticality of examining the numerous heating and cooling systems

available and their performance in a wide variety of structures, it is clear

that the present results should not be applied directly to specific situations,

but used as general guidelines. These guidelines can aid in retrofitting and

installing new equipment by reducing the many options from which to choose.

The dynamics of controls can also have a significant impact on the total

building energy consumption. In the present study, the dynamic minute-byminute interaction of air handling system controls on building zone loads and

zone air temperature was not investigated.

1 Douglas C. Hittle, The Building Loads Analysis System Thermodynamics (BLAST)

Program: Report Number CERL-TR-E-1 53 , Construction Engineering Research

Laboratory, June 1979.

2
3. SIMULATION PROCEDURE

3.1 THE BLAST PROGRAM

The Building Loads Analysis System Thermodynamics (BLAST) program was used to

predict the heating and cooling loads of a small office building and the annual

energy consumption of selected systems and strategies. The program is divided

into several distinct subprograms of which the building loads and air handling

system subprograms were used in this study. The building loads subprogram

requires a wide variety of input data including hourly weather data, thermophysical properties of the various surfaces as well as hourly internal heat

gain schedules generated by occupants, lights and electrical equipment. The

building is subdivided into thermally uniform zones and the hourly heating and

cooling loads and air temperatures in each zone are calculated. In order for

the calculation to proceed, hourly zone air temperature control profiles must

be specified. The zone air temperature control profiles are used to simulate

actual control operation within a building zone. The type of controls and the

air handling system capability to maintain desired conditions within the zone must be anticipated by the user of the BLAST program.

The hourly heating and cooling loads, zone air temperatures, and return air

temperature heat gains computed by the building loads subprogram are used along

with hourly weather data in the air handling system subprogram. The air handling system component characteristics and the control strategies are input by

the user. Peak heating and cooling equipment capacities, required to meet zone

loads, are calculated through preliminary simulations. Hourly energy consumption of the system components are calculated and monthly and annual values

reported

From the above description of the BLAST program, it is evident that the air

handling system subprogram does not "feed back" to the loads portion of the

program. In actual systems the feed back loop would tend to alter the zone

loads and zone air temperatures as well as induce changes in the operation of

the air handling system. The BLAST program is therefore in a sense a "quasi

steady-state" program, incorporating what has commonly been presumed first

order effects. Other research is currently underway to determine the effect on energy consumption of the dynamic interactions (minute-by-minute) of HVAC

control systems.

3.2 SMALL OFFICE BUILDING

A schematic of the small office building used in this study is provided in

figure 1. It is a three story lightweight structure measuring 30.48 m (100 ft)

per side. The external wall consist of a metal curtain on the outside, an airspace, 5.1 cm (2 in) of insulation, and 1.3 cm (0.5 in) of gypsum board on the

inside. Insulating glass is used, consisting of two 0.32 cm (1/8 in.) sheets

separated by an airspace. The glass to wall area is 31 percent. Ten thermal

zones are defined for the building and are shown in figure 1. Zone 1 consists

of the interior of the first and second floors, and zones 2, 3, 4, 5 each com- prise the first two floor's perimeter zones. The top floor's interior zone is

labeled zone 6 and the top floor is perimeter zones are labeled 7 through 10.

3
The hourly internal heat gain schedules due to occupants, lighting, and office

electrical equipment are presented in figure 2. Thirty percent of the heat

gain due to lighting is directed to the return air ducts. During normal

working hours, the number of occupants, the extent of lighting and electrical

usage are in the range commonly used for design purposes2>3. The nighttime

and weekend schedules are adjusted for non-occupancy. No infiltration is con- sidered during the occupied periods as it was assumed that the supply air fans

pressurized the building. Likewise, infiltration is not considered during the

unoccupied periods since the strategies investigated herein included intermittent or continuous operation of the supply air fans.

No alterations were made to the building envelope or occupancy schedules

throughout the simulations. Changes made to the exterior envelope to conform

to recommend standards^ for various climates were not considered because they

would tend to mask differences in system performance with climate. Also, such

changes were considered artificial since different building designs (not con- sidered here) would result in at least as great a variance in building heating

and cooling loads.

3 . 3 WEATHER DATA

The performance of HVAC systems is dependent on the weather either directly as

in the performance of economizer cycles or indirectly as it effects zone heating

and cooling loads. Typical Meteorological Year (TMY) and SOLMET^ hourly weather

data were used in the simulations. The weather data were used for seven locations: Washington, DC; Nashville, TN; Madison, WI; Seattle, WA; Dodge City, KS;

Lake Charles, LA; and Santa Maria, CA. The climates represented by these cities

are wide ranging. Tables 1 lists the heating and cooling degree days which

varied from 1498 to 7730 and from 84 to 2739, respectively.

2 Energy Conservation with Comfort, the Honeywell Energy Conserver's Manual

and Workbook, Honeywell, 1979.

3 Reference Building Descriptions, DoC Ref. l.b.9, July 17, 1980, Draft Report.

^ ASHRAE Standard 90-75, "Energy Conservation in New Building Design," American

Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc., 1975.

3 Hourly Solar Radiation-Surface Meterological Observations, TD-9724, National

Oceanic and Atmospheric Administration, Environmental Data Service, August

1978.

4
4. SIMULATION OF HVAC SYSTEMS AND STRATEGIES

Three systems commonly used in small office buildings were selected for

investigation. The three systems are:

1. Constant volume terminal reheat system serving the entire building,

2. Two independently-operated constant volume terminal reheat systems

serving separate zones of the building,

3. Variable-air volume terminal reheat system serving the entire building.

The first two systems have been popular in the past and are attractive on a first-cost bases. The more costly variable-air volume system has recently gain

reacceptance due to improved control and sensor technology and because of its

more efficient operation.

4.1 SPACE CONDITIONING

The air handling system capacities were sized such that during normal working

hours, the zone dry-bulb temperatures did not exceed 25.6 °C (78°F). Zone dewpoint temperatures were not allowed to exceed 18.3°C (65°F). The minimum zone air temperatures during occupied hours was 18.9°C (66 °F). These comfort limits

during occupied hours fall within the ranges established by the Federal

Government’s Emergency Temperature Building Restrictions and the ASHRAE

Standard 90-75; Energy Conservation in New Building Design.

In strategies employing temperatures setback during unoccupied hours, heating

is applied at 13.9°C (57°F) and zone air temperatures were allowed to fall to

12.8°C (55°F). These setback temperatures are those which are commonly

employed, and should not be considered as optimum values. Since it was felt

that, for the systems investigated, the BLAST program could not realistically

simulate temperature set-up for cooling, the zone air temperatures were allowed

to float during unoccupied hours in order to prevent the air handling system

from operating. Otherwise, the simulation program called for a simultaneous

cooling and reheating of supply air to exactly meet the pre-determined zone cooling loads.

Minimum outdoor air quantities were specified as a percentage of the total

system supply air. For the constant volume terminal reheat systems the minimum

quantity used was approximately .008 m-Vs (17 cfm) per person. When supply air

quantities were changed, the percent outdoor air was changed in order to maintain the same amount of minimum outdoor air. For the variable air volume

system, the minimum outdoor air quantities were approximately the same as for

the constant volume system when operating at peak block cooling loads. During

peak block heating, the minimum outdoor air quantity for the variable volume

system was reduced to approximately (.0028 m-Vs) (6 cfm) per person when no economy cycles were used.

5
4.2 CONSTANT VOLUME TERMINAL REHEAT SYSTEM SERVING THE ENTIRE BUILDING

The constant volume system consisted of a preheat coil, cooling coil and supply

fan which distributed the conditioned air to each zone. Heating and cooling

capacity modulation was accomplished by hot water reheat coils in the supply

air ducts leading to each zone.

This system is hereinafter referred to as the CV system. Thirty percent of the

lighting heat gain in the model was directed to the return air ducts. Due to

the small size of the building no return fan was considered. A portion of the

combined return air from all the zones was exhausted and an equal amount of

outside air introduced. All the strategies employed indicated that preheating

to the extent needed to prevent the cooling coil from freezing was negligible,

and therefore no preheat coil was considered.

The static pressure rise across the supply fan can have a large effect on fan

energy consumption. The amount of rise is dependent on the size and design of

the ductwork. A static pressure rise of 872 Pa (3.5 in. of water) was selected

as a typical value for this type of system.

4.2.1 CV System Strategies

Case A-l (Base Case)

The base case strategy was as follows

1. The system operated during working hours from 7 a.m. to 7 p.m., weekdays

within the desired comfort range. This included one hour prior to occu- pancy to establish habitable conditions. Zone air temperatures were allowed to range from 18.9°C (66°F) to 25.6°C (78°F). Zone reheat coils were activated when the zone air temperatures fell below 21.7°C (71 °F).

Since zone 1 experienced a relatively constant cooling load, no reheat

capability was added to this zone.

2. During the remaining hours, 7 p.m. to 7 a.m., and on weekends, the cooling

coil was turned off and zone air temperatures allowed to float. Reheat

coils were activated at zone air terapeatures of 13.9°C (57°F), reaching

their maximum heating capacity at zone temperatures of 12.8°C (55°F).

3. The cold deck supply air temperature during working hours fixed at 16.7°C

(62°F). A 1.7°C (3°F) cold deck throttling range permitted the supply air

temperatures to reach a minimum of 15 °C (59°F) when the cooling coil

experiences a small or near zero load. The cold deck temperature used was

set at the maximum allowable temperature that would prevent the interior

zone (which had the lowest sensible heat ratio) from exceeding a dew point

temperature of 18.3°C (65°F).

4. During the working hours the minimum outdoor air fraction was fixed at .10

of the total supply air quality at all otheer times no outside air was introduced

6
Case A-2

This strategy was the same as case A-l with an enthalpy economy cycle

added. In this cycle outdoor air quantity is proportioned with the return

air such that the enthalpy of the mixture entering the cooling coil

results in the smallest cooling coil load. The outdoor air quantity is,

however, limited to the minimum required for ventilation as previously

discussed

Case A

This strategy was the same as case A-l with a temperature economy cycle

added. In this cycle outdoor air is introduced (above the minimum needed

to satisfy ventilation requirements) only when the air temperature is

equal to or below the supply air temperature.

Case B-l

The following modifications to the base case strategy A-l were made.

1. The cold deck temperature was reset by the zone with the greatest

sensible cooling demand. The cold deck temperature reset range

was from 16.7°C (62°F) to 12.8°C (55°F).

2. The supply air quantities in all but the lower interior zone (zone 1, refer to figure 1) were adjusted to satisfy the peak

sensible cooling loads at a cold deck temperature of 12.8°C

(55°F). The lower interior zone supply air flowrate was sized

to meet the cooling load at a cold deck temperature of 16.7°C

(62°F). Since the lower interior zone experienced a relatively

constant cooling load, the effect of oversizing for this zone allowed the cold deck temperature to be controlled and reset by

the remaining zones.

3. Reheat capability was added to zone 1. This was necessary since

the lower cold deck temperatures expected during the summer months resulted in oversized cooling capacity and overcooling of

the space. Reheat coils in all zones were activated when zone air temperatures fell below 21.7°C (71 °F).

4. A fixed outdoor air ratio of .15 was used during normal working

hours. Due to the lower supply air quantities used in this

strategy, the resulting outdoor air quantities introduced into

the building were nearly the same as in the previously discussed

strategies. During unoccupied hours, no outside air was introduced

The intent of this strategy was to reduce the amount of reheat required to

satisfy the space conditioning requirements of the exterior and upper

interior zones. In addition, the supply fan energy consumption was reduced due to the lower average supply air flowrates.

7
Case B-2

This strategy was the same as case B-l , but modified by the addition of an enthalpy economy cycle.

Case C-l

This strategy was similar to case B-l except that the cold deck

temperature was reset by the outdoor temperature. Several simulations were required to establish a satisfactory reset schedule which would

maintain space conditions within the comfort range during occupied hours.

The reset control was 12.8°C (55°F) cold deck temperature at an outdoor

temperature of 21.1°C (70°F), increasing proportionally with decreasing

outdoor temperature of 16.7°C (62°F) at an outdoor temperature of 4.4°C

(40 °F)

Case C-2

This strategy was the same as case C-l modified to include an enthalpy

cycle.

Case 0

This strategy was the same as base case A-l except that the system

remained on continuously to maintain comfort. No outside air was introduced during the normally unoccupied hours.

4.3 SYSTEM INCORPORATING TWO CONSTANT VOLUME TERMINAL REHEAT UNITS

Many factors can influence the selection of zones in a building which should be

controlled by one air handling system. One such consideration is the reduction

in operating costs due to energy savings. The building envelope and the individual zone heating and cooling loads will vary from building to building so that zone optimization will also vary. Factors other than energy savings may

decide which zones should be controlled by a system. This is especially true

when selected zones are required to have continuous year round conditioning.

Two separate systems such as the ones utilized in the simulation may be attractive to potential owners and operators in that it can offer improved flexibility, energy savings and possibly lower initial cost than more complex systems

(such as VAV) serving the building.

The selection of zones to be controlled by one system was based on analysis of

the zone cooling loads encounteered for Washington, DC climate. The lower

interior (zone 1) and south-facing (zones 2, 7) zones were observed to have

high year-round cooling loads and were controlled by one system. The remaining

zones (zones 3, 4, 5, 6, 8, 9, 10) were observed to have highly variable cooling

and heating loads and were controlled by a second system. Thus, the second

system provides the opportunity for a higher cold deck temperature. This in

turn could result in reduced energy consumption due to less required reheat.

8
A static pressure rise of 623 Pa (2.5 in. of water) was selected for each air

handling system. Hereinafter, this system is referred to as a CV-CV System.

4.3.1 CV-CV System Strategies

In describing the following strategies, the air handling system controlling

zones 1, 2 and 7 is labeled CV-1 and the system controlling the remaining zones

of the building is labeled CV-2.

Case A-l (Base Case)

1. Both CV-1 and CV-2 systems operated weekdays during working hours from

7 a.m. to 7 p.m. within the comfort range. This included one hour

prior to occupancy to establish comfort conditions. Zone air temperatures were allowed to range from 18.9°C (66°F) to 25.6°C (78°F). Zone

reheat coils were activated when zone air temperatures fell below

21 .1 °C (71 °F)

As in the previous constant volume case, no reheat coils were applied

to zone 1. For Santa Maria, reheat was also not required for zone 2.

2. During the remaining hours and on weekends, the cooling coil was turned off and zone air temperatures allowed to float. Reheat coils were activated, however, when zone air temperatures fell below 13.9°C

(57 °F)

3. The cold deck supply air temperature during working hours was fixed at

17.5°C (63.5°F) for CV-2 and 16.7°C (62°F) for CV-1. Both cold decks

had throttling ranges of 1.7°C (3°F). These supply air temperatures

were the maximum that could be achieved and still satisfy comfort

space humidity conditions during the spring, summer, and fall months.

Maximum supply air temperatures were selected as it was felt that most owners and operators would want to optimize chiller efficiency.

4. During working hours a fixed minimum outdoor air ratios of .10 for

CV-1 and .09 for CV-2 is used. The resultant outdoor air quantities

are the same as the previous system, corresponding to approximately

.008 m^/s (17 cfm) per person. During unoccupied hours, no outside

air was introduced.

Case A-2

This strategy was the same as case A-l with an enthalpy economy cycle

added

Case A

This strategy was the same as case A-l with a temperature economy cycle

added

9
Case B-l

During the winter months the cold deck temperature for CV-2 was reset by

the zone of greatest demand. The cold deck temperature was allowed to

vary from a minimum of 17. 5° C (63.5°F) to a maximum of 21.1°C (70°F). The

winter months were defined as those months for which the outdoor temperature during the daylight hours 9 through 16 did not exceed 16. 7° C (62°F)

more than 10 percent of the time*.

Using the above definition for winter/summer conditions, it was not

possible to implement strategies B-l, B-2 , B-3, for Lake Charles and

Santa Maria. Except for this modification, all other parameters remained

the same as the base case.

Case B-2

Same strategy in case B-l, except with an enthalpy economy cycle added.

The economy cycle here and in succeeding strategies were applied year

round

Case B-3

Same strategy as case B-l, except with a temperature economy cycle added.

Case C-l

During winter months the cold deck temperature for CV-2 was reset as a function of the outdoor temperature*. The cold deck temperature was reset

as follows: 17.5°C (63.5°F) at an outdoor temperature of 12.8°C (55°F)

increasing to a maximum value of 21.1°C (70°F) at an outdoor temperature

of 10° C (50°F)**. All other parameters remained the same as in the case

B-l.

Case C-2

Same strategy as case C-l, except with an enthalpy economy cycle added

Case C-3

Same strategy as case C-l, except with a temperature economy cycle added

* Derived from data supplied by Engineering Weather Data; Dept, of the Air

Force, the Army, and the Navy, AFM 88-29, TM 50785, NAVFAC P-89, 1978.

** As in strategies B-l, B-2 and B-3 the definition used for winter months did

not permit implementation of strategies C-l, C-2 and C-3 for Lake Charles

and Santa Maria.

10
Case D

Both air handling systems operated continuously to maintain com ort

conditions. No outdoor air was introduced during the unoccupied periods.

All other conditions remained the same as the base case A-l

4.4 VARIABLE-AIR VOLUME TERMINAL REHEAT SYSTEM

The variable-air volume system served all ten zones of the small office

building. Zone supply air quantity was modulated in response to the zone heating or cooling requirements. Zone thermostats controlled supply air

dampers from fully open at 25.6°C (78°F) to a selected minimum value at and

below 24 ,4°C (76°F).

A supply fan with variable pitch inlet vanes was simulated with a fan

efficiency of 65 percent in contrast to the 70 percent efficiencies used in

the constant volume systems. Static pressure rise the fan at peak air flow

conditions was indentical to the single constant volume system (872 Pa (3.5 in.

H2 O)). A description of the various strategies simulated for the VAV system

follows

4.4.1 VAV System Strategies

Case A-l (Base Case)

1. The system operated during working hours, 7 a.m. to 7.p.m. as the

previous systems base case strategies already described. During

these hours zone air temperatures ranged from 18.9°C (66°F) to 25.6°C

(78°F). Zone reheat coils were turned on when zone air temperatures

fell below 21.7°C (71°F). As in the other systems, zone 1 contained

no reheat coils and in the case of Santa Maria, zone 2 reheat was

turned off.

2. During the remaining hours and on weekends, the cooling coil was turned off and zone air temperatures allowed to float. Reheat coils

were, however, activated when zone air temperatures fell below 13.9°C

(57 °F)

3. The cold deck supply air temperature was fixed at 12.8°C (55°F) with a 1.7°C (3°F) throttling range. The cold deck was set at this value to

insure adequate dehumidification during mild weather when zone supply

air quantities may be reduced.

4. A minimum outdoor air fraction was fixed at .15 during working hours.

The resultant outdoor air quantity was comparable to the two previous

systems when the building experienced peak block cooling loads. During

peak block heating and resultant air quantity was sufficient to meet a minimum recommended value of .0028 m^/s (6 cfm) per person. During

unoccupied hours no outside air was introduced.

11
5. The variable air volume terminals had a minimum fraction of 0.2. As

previously stated, the maximum supply air was available at zone air

temperatures of 25.6°C (78°F), falling to 20 percent of the maximum

value at zone air temperatures of 24.4°C (76°F) and less.

Case A-2

This strategy was the same as the base case strategy, case A-l , modified

by the addition of an enthalpy economy cycle.

Case A

This strategy was the same as base case A-l modified by the addition of a temperature economy cycle.

Case B-l

The base case strategy, case A-l, was modified as follows:

1. The cold deck temperature was reset by the zone requiring the most

cooling. The cold deck temperature reset limits were 12.8°C (55°F) to

16.7 °C ( 62 °F)

2. The supply air quality for the lower interior zone (zone 1) was readjusted to satisfy the cooling load at a cold deck temperature of

16.7°C (62°F) when the zone air damper was fully open. Due to the

constant cooling load experienced by the lower interior zone, this

adjustment was necessary in order to prevent the zone from controlling

the cold deck temperature at the low value of 12.8°C (55°F). During

the winter months when the cold deck temperature was expected to be in

the neighborhood of 16.7°C (62°F), the interior zone supply damper

would be fully open, whereas during the summer months the damper could

modulate the supply air sufficiently to prevent overcooling. As a result, it was determined that the interior zone required no reheat.

Case B-2

This strategy was the same as case B-l , except with an enthalpy economy

cycle added.

Case B-3

This strategy was the same as case B-l, except with a temperature economy

cycle added.

Case C—

This strategy was similar to case B-l strategy except that the cold deck

temperature was reset by the outdoor temperature. The reset temperatures

were as follows: 12.8°C (55°F) at an outdoor temperature of 21.1°C (70°F)

increasing proportionally with decreasing outdoor temperature to 16.7°C

(62°F) at an outdoor temperature of 4.4°C (40°F).

12
Case C-2

This strategy was the same as case C-l strategy with an enthalpy economy

cycle added.

Case C-3

This strategy was the same as case C-l strategy with a temperature economy

cycle added.

Case D

Comfort conditions were maintained continuously throughout the day and

night all other conditions remain the same as the base case ( A— 1 ) strategy

and no outside air was introduced into the space during the unoccupied

hours

13
5. RESULTS AND DISCUSSION

Annual energy consumption results are presented in tables 2 through 8 for the

three HVAC systems and the seven cities discussed previously. Each table lists

hot water, chilled water and fan annual energy consumption in the upper portion

of each block for each of the three HVAC systems. The first column in each

table briefly describes the strategy employed for the systems. For a selected

strategy, comparisons between HVAC systems can be made for the annual energy

consumption for chilled water, hot water and fan energy consumption. The

numbers appearing in the lower portion of each block represent the ratios of

energy consumption relative to the base case strategy (case A-l) for that particular system . In the ensuing discussions, that following abbreviations are used to describe each HVAC system:

CV system: constant volume terminal reheat system serving the entire

building,

CV-CV system: two independently operated CV systems; one serving zones 1, 2,

and 7 and the other serving the remaining zones.

VAV system: variable air volume terminal reheat system serving the entire

building.

Each strategy presented in tables 2 through 8 differ in some respects for each

of the three systems due to constraints imposes by the system. For example, in

the base case strategy (case A-l) the cold deck supply air temperatures for the

CV system was 16.7°C (62°F); for the CV-CV system they were 16.7°C (62°F) for

the first air handling unit and 17.5°C (63.5°F) for the second; for the VAV

system the cold-deck temperature was 12.8°C (55°F). The differences in some parameters for the same basic strategy in each of the three systems and the reasons for selection were discussed in the previous section.

5.1 COMPARISON OF SYSTEMS - (tables 2 through 8)

The results of tables 2 through 8 indicate that for each of the different

strategies the VAV system is the most efficient system (least fan, chilled

water and hot water annual energy consumption) for all seven cities.

In comparing the base case strategy (A-l) for the VAV system with strategy B-2

(zone controlled cold deck with enthalpy economy cycle) for the CV-CV system,

it is found that the former system uses less hot water and fan energy in all seven cities. It also uses less chilled water in all cities, except Madison

and Seattle. Since for each strategy the VAV system generally outperforms the

other two systems, the ensuing discussion centers on comparison of the CV-CV

and CV systems.

In comparing the CV and CV-CV system for the cases A-l through A-3 , the hot

water and chilled water energy consumption are approximately the same. This is

expected since in these cases the CV-CV is operating as if it were a single

system (CV system). Comparisons in fan energy consumption for these two systems should not be made, since in each system differences in duct design and

14
installation could alter the relative values. (In the present study, a static

pressure rise of 623 Pa (2.5 in H2 O) was used for each of the supply fans in

the CV-CV system whereas a value of 872 Pa (3.5 in 11? 0) was used for the CV

system.

)

In comparing zone-controlled cold deck strategy (case B-l ) for the CV and CV-CV

system, both systems use approximately the same energy for chilled water. However, the percent savings in hot water energy consumption (reheat energy) for

the CV-CV system over the CV system range from 21 percent to 35 percent,

increasing with decreasing number of cooling degree days in each city. The

large consumption in hot water for the zone-control CV system is due to the

fact that the interior zone requires reheating to prevent overcooling of the

space when the cold deck supply air temperature is low. For strategy B-2 , the

percent savings in hot water of the CV-CV system over the CV system is approximately the same as case B-l when an enthalpy economy cycle is added.

In the zone-controlled and outdoor air-controlled cold deck strategies (case B

and C) , Lake Charles and Santa Maria are not included, since cold deck reset

was applied only during the winter months and thus did not apply for these

cities. As discussed in section 4.3.1, the winter months were defined as those

months for which the outdoor temperature during the daylight hours 9 through 16

did not exceed 16.7 °C (62 °F) more than 10 percent of the time.

As mentioned previously, it might be possible to further improve the performance

of the CV-CV system by optimizing the zoning of each system according to the

climate. In addition, variations in performance will result if the zone- controlled strategy (case B) is applied year round. In such a case deck supply

air temperature would have to be in a lower range, (for example, 12.8°C (55°F)

to 1 7 . 5 °C (63.5°F) instead of 17.5°C (63.5°F) to 21.1°C (70°F)). These

possibilities were, however, not investigated.

Increased savings in chilled water consumption is achieved by the CV-CV system

over the CV system for case B-2. This is accomplished by allowing the economy

cycle to offset cooling coil energy consumption at higher outdoor temperatures.

This was possible since the cold deck temperature for one air handling unit in

the CV-CV system was reset to 21.1 °C (70°F), whereas for the CV system (as well

as the second air handling unit on the CV-CV system) , the allowable maximum

cold deck temperature was 16.7°C (62°F).

Hot water consumption is also less for the CV-CV system as compared to the CV

system for strategy C-l (cold deck temperature reset by outdoor air temperature).

The percent savings in. hot water generally follow the same trend as in the zone- controlled strategies (cases B-l, B-2, B-3), however, they average approximately

10 percent more in energy savings. Savings in chilled water energy consumption

are also realized and range from 5 percent to 13 percent. The results for cases C-l and C-2 should be regarded as approximate since in some cities the

simulations indicated variations in the number of hours for which there were

unmet cooling loads. An unmet cooling load is defined as the additional cooling

required by a zone which cannot be met by the zone air handling system. In the

worst case, the maximum percentage of unmet cooling loads in any zone did not

exceed 6 percent, with the remaining zones having a much lower percentage.

15
COMPARISON OP STK ATF.O I PS ( tables 2 through 8)

\ . Hnr.n Case Strategy ( A-l ) and Base Case with Enthalpy Economy (A-?.)

In comparing st: ra top. ies A-2 and A-l with (heir respective base cast* strategies

(A-l) for the CV-CV system and CV system, neg.l igible differences are apparent

in the hot and chilled water and fan cnerpv consumption. Therefore, in the

ensuing discussion of strategies A-2 and A-3 , comparisons are made between the

CV and VAV system with the results for the CV system also applying to the CV-CV

svst i>m

Pan E nergy Consunpt i on : There are negligible differences in fan energy

consumption for strategy A-2 compared to strategy A-l for all systems and for

all seven cities.

Hot Water Energy Consumption: There is an increase in hot water energy

consumption for the CV system over the base case ranging from 3 percent to 7 percent for the five cities. Seattle and Santa Maria indicated a higher

percentage increase of approximately 12 percent. The VAV system indicates a much smaller increase in hot water energy consumption, ranging from 1 percent

( <i 2 percent for all seven cities.

The increase is due primarily to the throttling of the cooling coil to slightly

lower cold deck temperatures (the maximum throttling was 1.8°C (3°F)), which,

it not modulated by reheat would cause overcooling of zones. This is especially

true in mild climates such as Seattle and Santa Maria where more hours are available for operation with an economy cycle.

The VAV system exhibits a smaller increase in the amount of reheat energy

consumption than the MV system when both systems include enthalpy economy

cycles. Both systems have approximately the same size cooling coil since at

design conditions they both must meet similar block (total) cooling loads. As

the results indicate, much less chilled water is used by the VAV system, implying

that tiu.' cooling coil for this system is operating most of the time at part

load conditions. Since the cooling coil for the VAV system is throttled back

most of the time, any decrease in the cooling coil load (by the addition of an economizer) will not substantially Increase the throttling.

Chilled Water Energy Consumption; There is a wide variation in savings in

chilled water energy consumption by employing an enthalpy economy cycle, as shown in the following table:

hake Washington Santa

Charles Nashv i lie DC Dodge Mad i son Seat t ie Ma r i a CV

and

CV-CV

m 36% 44% 41% 51% 75% 60%

VAV 10% 20% 27% 25% 33% 55% 38%

lb
where the percent savings are with respect to the base case strategy A-l. The

savings, as expected, generally increase with decreases in the number of cooling

degree days (see Table 1). The VAV system shows less savings with respect to

its base case strategy, since with less the supply and return air flow rates

are reduced in mild and cold weather, with less reheating being applied than

the CV system, and the cooling coil load is small. The CV system, for the same outdoor conditions, will have a larger cooling coil load and introducing outside

air to reduce the cooling coil load will therefore have a much greater effect

on the CV system than the VAV system.

B. Temperature (case A-3) and Enthalpy (case A-2) Economy Cycles Added to the

Base Case Strategy

Fan Energy : As in the previous case there are negligible differences in fan

energy consumption for case A-3 (temperature economy cycle) as compared to case

A-2 (enthalpy economy cycle) for all systems in the seven cities.

Hot Water Energy Consumption : For the CV system there is a one to two percent

savings in hot water energy consumption for case A-3 as compared to case A-2 in

the five cities, although there is still an increase in hot water energy

consumption with respect to the base case, A-l. Savings for Seattle and Santa

Maria range slightly higher, from two to five percent, relative to case A-2.

The VAV system, in general, indicates negligible differences between strategies

A-2 and A-3.

Chilled Water Energy : The following table lists the percentage savings in

chilled water energy consumption by using an enthalpy economy cycle (case A-2)

as compared to a temperature economy cycle (case A-3), for the CV and VAV

systems

Lake Washington Santa

Charles Nashville DC Dodge Madison Seattle Maria

CV 7% 11% 11% 11% 16% 39% 40%

VAV 8% 10% 13% 11% 13% 36% 29%

As the above table indicates, the percentage savings chilled water energy

consumption for case A-2 as compared to case A-3 generally increases with a decrease in the number of cooling degree days. The increases are primarily due

to two effects. One effect in using a temperature economy cycle is that outside air is only introduced when the outside air temperature is less than or equal to the cold deck temperature. For an enthalpy economy cycle (as well as return air economy cycle) , outside air is introduced when its enthalpy is less

than or equal to the return air enthalpy. Thus, outside air in the enthalpy

economy cycle reduces the cooling coil load at higher outdoor temperatures

(above the cold deck temperature) than the temperature economy cycle. The

dramatic savings in chilled water consumption for Seattle and Santa Maria as compared to the other five cities listed in the table can be understood with

17
the aid of the third column in table 1. These numbers represent the percentage

of temperature bin hours encountered between outdoor temperatures ranging from

13.9°C (57°F) to 22.2°C (72°F) for the daylight hours from 9 a.m. to 4 p.ra. The outdoor temperature of 13.9°C (57°F) represents an approximate value above

which a temperature economy cycle (for cold deck temperatures set at 13.9°C

(57 °F)) is inoperable. The representative outdoor temperature of 22.2 °C (72 °F)

is used as a typical return air temperature for which an enthalpy economy cycle

operating at or below this outdoor temperature can partly offset the cooling

coil load. It is evident from table 1 that an enthalpy economy cycle is very

effective for Seattle and Santa Maria since a significant portion of the

operating hours are spent within this temperature range.

C. Base Case Strategy (A-l) and Zone-Controlled Cold Deck Strategy (B-l)

Fan Energy Consumption : For the CV system, fan energy consumption is lower for

the zone-controlled cold deck stratety (case B-l) as compared to the fixed cold

deck strategy (case A-l). The decreases in fan energy consumption range from

25 to 28 percent for the seven cities. The savings are realized since all but

the interior zone air flow rates are reduced for case B-l to meet peak cooling

demands at a 12.8°C (55°F) cold deck temperature. In the simulations, the

interior zone air flow rate remained the same as the base case (sized to meet

peak cooling loads at a 16.7°C (62°F)) in order to allow the reamining zones to

control the cold deck temperature.

No changes in fan energy consumption are encountered with the CV-CV system

since zone air flow rates remain the same as in the base case, A-l.

The VAV system shows a general increase in fan energy for zone-controlled cold

deck. For Nashville, Washington, DC, and Madison the percentage increase

ranges from 11 to 14 percent. The remaining cities show smaller increases,

averaging approximately 2 percent. The increase in fan energy consumption is

due primarily to the variable the air volume dampers opening to a larger

extent than for the base case. With a zone-controlled cold deck strategy,

cooling capacity is reduced sequentially by first increasing the cold deck

temperature. Since zone cooling capacity is reduced by this method, a reduction in zone supply air quantities is not needed. In fact, zone air quantities

generally increase with increasing cold deck temperature (for a given zone cooling load) until at least one of the 10 zones with the greatest cooling

demand has a maximum zone air damper opening. Thereafter, further reductions

in zone cooling capacities is accomplished by gradually closing the dampers.

Hot Water Energy Consumption . The following table is a summary of the percent

decrease in hot water energy consumption for zone-controlled cold deck (case

B-l) as compared to the fixed cold deck (case A-l).

18
Lake Washington Santa

Cha r 1 es Nashvil 1 e DC Dodge Madison Seat 1 1

Ma r 1 a

CV 0 % -14% -26% -22% -32% -22% -6%

cv-cv N/A 12% 16% 14% 16% 21% N/A

VAV 21% 31% 27% 9% 19% 13% 26%

For the CV systemi hot wate r energy consumption increases when employing a zone controlled cold deck. The reason for the i ncrease is due to the fact that the

interior zone requires the addit ion of a reheat coil in order to prevent

overcooling of the space when the cold deck temperature Is low. Ry oversizing

the cooling capacity in the Interior zone, the remaining zones can thereby

control the cold deck temperature. The results vary widely from city to city

since the increases in reheat energy needed to satisfy interior zone cooling

are offset by decreases in the required reheat energy in the remaining zones

(when the cold deck temperature is high).

Using a zone controlled cold deck strategy in the CV-CV and VAV systems results

in a higher percentage savings over the base case, A-l . For the CV-CV system

the savings range from 12 to 21 percent, generally increasing with a decrease

in the number of cooling degree days. The VAV system employing zone controlled

cold deck shows savings over the base case ranging from 9 to 31 percent.

The generally higher savings in hot water energy consumption for the VAV system

as compared to the CV-CV system is in part attributable to those periods when

mild weather conditions prevail. During these times a maximum supply air temperature is sufficient to meet the peak zone cooling loads and simultaneously

provide a minimum of reheat to those zones which require heating. The savings

do not appear to follow any correlation with degree days, as indicated in the

above table.

Chilled Water Energy Consumption . Employing zone controlled cold deck for the

CV system results in savings over the base case ranging from 7 percent for Lake

Charles to 19 percent for Seattle. The savings generally increase with a decrease in cooling degree days except for Santa Maria where the savings are approximately 11 percent.

Savings for the CV-CV system follow the same trend as the CV system, however,

the magnitudes are less; ranging from 6 percent for Nashville to 1 5 percent for

Seattle

The percent savings over the base case for the VAV system are generally smaller

than those for the CV-CV and CV system. The savings range from 1 to 8 percent

for 6 cities and is approximately 11 percent for Madison.

19
The lower percent savings for the VAV system as compared to the CV and CV-CV

systems may be understood in part by examining the cold deck reset control for

this system during different seasons for the year. During the summer as well

as mild weather periods the VAV system modulates zone cooling capacities by

determining the maximum permissable cold deck temperature which will cause the zone of greatest cooling demand to operate with the dampers fully open. During

these occasions the dampers in the remaining zones will be partially closed in

order to reduce cooling capacities to match zone cooling loads. Since the zone supply air dampers have zone operating range (capable of modulating zone supply

air flowrates by 80 perecnt) the zone cooling loads can also be met by maintaining a fixed cold deck temperature and modulating the dampers alone. In the

latter case the zone air dampers would open and close to a greater extent than

the former case to meet the same zone cooling loads. Thus with or without zone controlled cold deck during mild or summer weather, the cooling coil load will

be nearly the same, approximately the sum of the zone cooling loads and return

air heat gains. During the winter months, however, when most of the zones are heating and the remaining zones operating at less than maximum air flow rates,

savings are realized since the cold deck temperature is reset upwards and

supply air flow rates are lower.

The additional complexity of the zone controlled deck for the VAV system in

contrast to the CV and CV-CV zone controlled system is reflected in the lack of

a consistent trend in hot water and chilled water savings with cooling and

heating degree days. The amount of savings using zone controlled cold deck on VAV systems may be dependent on other meteorological factors such as solar

radiation, cloud cover, and daily means and extreme temperatures during the

summer, winter, and intermediate months. Other factors which will determine

savings are the minimum air flow ratio of the VAV system as well as the thermal

responsiveness of the various zones within the building. The latter factor is

of course applicable to the CV and CV-CV systems.

D. Temperature Economy Cycle (case B-3) and Enthalpy Economy Cycle (case B-2)

for Zone-Controlled Cold Deck Systems

Fan Energy Consumption : Tables 2 through 8 indicate negligible differences in

fan energy consumption between cases B-2 and B-3 for all systems in the seven cities

Hot Water Energy Consumption : For the CV system there is a one to two percent

savings in hot water energy consumption for strategy B-3 as compared to strategy

B-2 in the five cities. Savings relative to strategy B-2 for Seattle and Santa

Maria range slightly higher, from 2 to 5 percent. Savings for the CV-CV system

are nearly identical to the CV system results. The VAV system in general

indicates negligible differences between strategies B-2 and B-3.

Chilled Water Energy Consumption : For the CV system the percentage decrease in

chilled water consumption for zone-controlled cold deck systems with enthalpy

economy cycle (case B-2) as compared to zone-controlled cold deck with temperature economy cycle (case B-3) is the same (within 3 percent) for five cities as in the results stated for cases A-2 and A-3 (fixed cold deck temperatures).

20
The CV-CV system indicates nearly identical results as the CV system for five

cities. The results, however, are not reported for Lake Charles and Santa

Maria since in these cities it was not possible to use a zone controlled cold

deck temperature strategy.

Likewise, the VAV system shows similar percent decreases in chilled water

consumption as found in the comparison of cases A-2 and A-3.

E. Zone-Controlled (case B-l) and Outdoor Air-Controlled (case C-l) Cold Deck

Strategies

Fan Energy Consumption : For the CV-CV and CV system there are negligible

differences in fan energy consumption for the outdoor-air controlled cold deck

(case C-l) and the zone-controlled cold deck (case B-l) strategies. For the

VAV system savings in fan energy using outdoor-air controlled cold deck strategy as compared to the zone-controlled cold deck strategy range from 4 to 7 percent for Nashville, Washington, DC, Dodge, and Madison; the remaining cities

show negligible differences.

Hot Water Energy Consumption : The following table lists the approximate

decrease in hot water consumption by using zone-controlled cold deck temperature (case B-l) as compared to the outdoor air-controlled cold deck temperature

(case C-l).

Lake Washington Santa

Charles Nashville DC Dodge Madison Seattle Maria

CV 23% 22% 3% 14% 9% 13% 13%

CV-CV N/A 3% 1% 0% -2% -1% N/A

VAV 5% 7% 4% 4% 1% -14% 5%

As the

water

above table

consumption.

indicates

For the

, all systems

CV system, the

generally

savings

show a decrease in

range from 3 percent

hot

to 23

percent. The CV-CV system indicates negligible differences in hot water

consumption for all cities. Percent savings in hot water consumption for the

VAV system ranges from 1 percent to 7 percent for 6 cities. Seattle, however,

indicates an increase of 14 percent.

Chilled Water Energy Consumption : The following table lists the percent savings

in chilled water consumption using zone-controlled cold deck (case B-l) compared

to an outdoor air-controlled cold deck (case C-l).

21
Lake Washington Santa

Charles Nashville DC Dodge Madison Seattle Maria

cv 14% 14% 3% 11% 10% 11% 12%

CV-CV N/A 1% 0% 0% 0% -1% N/A

VAV 4% 1% 1% 1% 1% -7% 1%

For the CV system savings in chilled water consumption range from 10 percent

to 14 percent for 6 cities; Washington indicates a much lower savings of 3

percent

Negligible changes in chilled water savings is apparent for the CV-CV system,

as well as for the VAV system. Seattle, however, exhibits a similar anamolous

result in chilled water consumption as it does for hot water consumption. An

increase of 7 percent in chilled water energy consumption results when a zone controlled strategy is employed rather than an outdoor air-controlled strategy.

It is apparent that the comparative results in hot water and chilled water

energy consumption for the outdoor-air controlled strategies (case C-l ) with

the zone-controlled cold deck strategies (case B-l) can vary from city to city.

A number of effects can be identified which can cause these variations.

The percent changes in hot water and chilled water energy consumption using an outdoor temperature controlled cold deck over a zone controlled strategy can vary because the former strategy does not optimize according to zone and

building heating and cooling loads. In some instances the outdoor temperature

may dictate a cold deck temperature which is lower than that which would have

occurred if the system were zone controlled. In this instance, chilled water

consumption and hot water consumption (for zones which require heating) would

be higher than the zone-controlled cold deck strategy.

Zone cooling and heating requirements will depend not only on outside air

temperature but also on changing internal heat gains, solar heat gains, and how

responsive each zone is in translating these gains into zone loads. These

factors will also vary for different building designs, orientation, and occu- pancy (usage) schedules so that optimum outdoor air controlling cold deck

strategies require careful attention for each building and climate. These

temperatures must be selected to insure few hours for which there is undercooling yet high enough cold deck temperature and outdoor temperature as to be

effective

Finally, for buildings which are highly responsive to solar and internal loads,

outdoor cold deck temperature reset will not be as effective as applying the same strategy to a building which is more responsive to changes in the outdoor

temperature. Obviously, in the former building a low enough outdoor temperature must be selected to reset the cold deck temperature in order to meet the

22
large temperature-independent heat gains. In these cases zone controlling

would be expected to generate even higher energy savings.

Base Case Strategy (cases A-l) with Continuous Conditioning Strategy (case D)

From tables 2 through 8 it is apparent that the continuous conditioning

strategy (case D) compared to the base case strategy uses much more fan, hot

water, and chilled water energy. In case D strategy, the CV-CV system is

operating as if it were a single system so that the following discussion for

the CV system applies to the CV-CV system as well.

Fan Energy Consumption ; The ratio of fan energy consumption for case D

strategy compared to the base case A-l ranges from 1.5 to 2.5 for the seven

cities. The VAV system and the CV system have approximately the same ratio

for each city. The energy consumption for case D is less than the ratio of

total annual hours to occupied hours (approximately 2.8). This is due to the

fact that the fans operated intermittantly during unoccupied hours to satisfy

zone heating or cooling loads.

Hot Water Energy Consumption : The ratio of hot water energy consumption for

case D strategy compared to the base case for the CV system ranges from 3.5 to

4.0 for the seven cities. For the VAV system this ratio ranges from 2.8 to 4.2

for five cities; Lake Charles and Santa Maria having ratios of 5.9 to 8.3,

respectively.

The extreme high increases in reheat energy (above the ratio of annual hours to

occupied hours) is due primarily to reheating that is needed during the nor- mally off hours during the summer, as was observed in the monthly energy usage

summaries obtained in the simulations. During these periods the internal heat

gains are low so that only a small amount of zone cooling is required. In

order to achieve the low cooling capacities for the VAV and CV systems, the zone supply air temperature must be increased by reheating. Another contributing factor is the raised zone air temperatures during unoccupied periods for

winter heating.

Chilled Water Energy Consumption . Chilled water energy consumption ratios for

the CV system range from 2.2 to 2.7 for the seven cities. For the VAV system

the ratios are lower, ranging from 1.4 to 1.9. The lower ratios exhibited for

chilled water energy consumption as compared to hot water consumption is

primarily due to the small amount of cooling required during the unoccupied

hours. The VAV system exhibits a lower ratio than the CV system since zone cooling capacities can be lowered by reducing the zone supply air flow rates.

23
6 . SUMMARY

Using the RLAST computer program, a typical small office building was modeled

along with commonly employed systems and strategies. Annual energy consumption

results (at the systems level) are presented for hot water, chilled water and

circulating fans for seven geographical locations. The systems modeled were a variable air volume, (VAV) reheat system, a constant volume terminal reheat

(CV) system, and two independently operating constant volume terminal reheat

(CV-CV) systems.

For comparable control strategies, the variable air volume (VAV) system (with

a 0.2 minimum air volume ratio) exhibited the least energy consumption for

chilled water, hot water and circulating fan. The system incorporating two

independently operating constant volume terminal reheat units (CV-CV) ranked

second in energy consumption while the single constant volume terminal reheat

(CV) system ranked last. The results are summarized in the bar chart of

Figure 3 for the base case strategy, (case A-l). Moreover, the VAV system

with the least efficient strategy (excluding continuous operation) (case A-l)

generally used less chilled water and reheat hot water than the CV system with

a fixed cold deck and enthalpy economy cycle (case A-2) and the CV-CV system

with a zone-controlled cold deck and enthalpy economy cycle (case B-2). The

strategies for the latter two systems were judged the best of the ones investigated in this report and are presented in figure 4.

One of the more effective strategies investigated for reducing chilled water

energy consumption incorporated the use of economy cycles, particularly

enthalpy economy cycles. Figure 5 summarizes the relative effectiveness of

employing enthalpy economy cycles for the three systems in seven cities. For

each HVAC system, the percent savings in chilled water energy consumption for

a fixed cold deck strategy with an enthalpy economizer (case A-2) relative to

a fixed cold deck temperature alone (case A-l) was found to be approximately

equal to the percent savings in chilled water energy consumption for a zone- controlled cold deck strategy with enthalpy economizer (case B-2) relative to

a zone controlled cold deck temperature alone (case B-l )

The results described above were found to apply for the resulting increases in

hot water energy consumption when economy cycles are employed. Therefore, the

results presented in figure 5 represent the energy effectiveness of enthalpy

economy cycles when added to either a fixed cold deck strategy (case A-2 versus

case A-l) or a zone-controlled cold deck strategy (case B-2 versus B-l).

As indicated in figure 5, the percent savings in chilled water energy

consumption resulting from the addition of enthalpy economy cycles follow the same trends for all three HVAC systems in the seven cities. A discussion of

these trends with locality was made in Section 5.

Referring to figure 5, the VAV system indicates the smallest chilled water

savings when enthalpy economy cycles are added, ranging from 3 percent to 55

percent depending on locality. The chilled water savings for the CV and CV-CV

systems are approximately the same in magnitudes and trends. The smallest

savings (approximately 18 percent) are for Lake Charles and the largest are for

Seattle and Santa Maria (55 percent to 76 percent).

24
Figure 5 also indicates that for all three systems, the percent increases in

hot water energy consumption are small when enthalpy economy cycles are employed. The VAV system indicates a maximum increase of approximately 3 percent whereas for the CV and CV-CV systems the increases are slightly higher;

ranging from 2 percent to 12 percent. As discussed in Section 5 these

increases are primarily due to the automatic throttling back (by as much as 1.7°C (3°F)) of the cold deck supply air temperature when outside air is used

to offset the cooling coil load.

Figure 6 show results for percent changes in hot and chilled water energy

consumption when a temperature economizer is added to a zone-controlled cold

deck (case B-3 relative to case B-l ) or added to a fixed cold deck (case A-3

relative to case A-l). In figure 6, the trends in chilled water savings with

locality using temperature economy cycles are very similar to those found in

using enthalpy economy cycles (figure 5) however; the magnitudes of the savings

are reduced by approximately 10 percent to 30 percent, with the largest reductions occurring for Seattle and Santa Maria. As discussed in Section 5, the

larger difference in chilled water savings for these two cities are attributable to the greater number of hours for which the outdoor temperature ranges

between the cold deck temperature and return air temperature. For this outdoor

temperature range, a temperature economy cycle is relatively ineffective as compared to enthalpy or return air economy cycles.

Figure 6 indicates a small increase in hot water energy consumption when

temperature economy cycles are added. The VAV system indicates the smallest

increase (approximately 1 to 3 percent) of the three systems. The remaining

two systems show increases ranging from approximately 3 percent to 7 percent.

For the small office building model, adding a zone-controlled cold deck strategy

alone (case B-l) in place of a fixed cold deck strategy (case A-l) resulted in

chilled water energy savings ranging from 7 percent to 11 percent. Considering

this strategy as a retrofit option where the ductwork is unchanged, fan energy

consumption was reduced by approximately 27 percent. The fan and chilled water

energy savings were offset by increased reheat hot water energy consumption,

averaging 17 percent for the seven cities.

The results obtained by the above strategy (see discussion in Section 5) were

due to reduced supply air quantities in the perimeter zones (sized to satisfy

the peak perimeter zone cooling requirement at the lowered cold deck temperature

of 12.8°C (55°F) in place of 16.7°C (62°F)), as well as the necessity for adding

reheat coils to the interior zone to prevent overcooling.

When enthalpy or temperature economy cycles are used, replacing a fixed cold

deck strategy (cases A-2 , A-3) with a zone-controlled cold deck strategy (cases

B-2, B-3) results in an average 4 percent increase in chilled water energy

consumption. This negative result is probably due to the addition of reheat in

the interior zone for the zone-controlled cold deck strategy, resulting in

higher cooling loads during the summer months (when the interior zone reheat is

expected to operate). Economy cycles during the summer months would not be

expected to be effective in reducing the cooling coil load.

25
Of the strategies investigated for the CV system in the small office building

model, the ones which appear to have the clearest impact on overall energy

savings are the fixed cold deck with either enthalpy or temperature economy

cycles (case A-2 or A-3)

In contrast to the mixed results obtained with the CV system, zone-controlled

cold deck strategies for the VAV and CV-CV systems indicate definite savings in

energy consumption. A bar chart is presented in figure 7a for the relative

savings in chilled water and hot water energy consumption for the VAV and the

CV-CV systems when a fixed cold deck temperature strategy (case A-l) is

replaced by a zone-controlled cold deck strategy (case B-l). Results for the

CV-CV system for Lake Charles or Santa Maria are not shown since their climates

would result in uncomfortably high humidities within the building.

As shown in figure 7a, modest savings in chilled water energy consumption

ranging from 1 percent to 11 percent are obtained for the VAV system. For the

CV-CV system, chilled water savings range from 5 percent to 14 percent.

Larger savings in reheat hot water energy consumption are obtained, ranging

from 9 percent to 30 percent for the VAV system and from 12 percent to 21

percent for the CV-CV system.

Fan energy consumption for the VAV system increases—ranging from 1 to 14

percent for the seven cities. As discussed in Section 5, the increase is

expected since the supply air damper opening on some zones must increase to

meet the zone cooling load at the higher cold deck supply air temperatures.

Fan energy consumption for the CV-CV system did not change since supply air

flowrate remained the same as with the fixed cold deck strategy.

Figure 7b is a bar chart for chilled water and hot water energy savings when

zone-controlled cold deck is added to the VAV and CV-CV systems having preexisting enthalpy economizers (case B-2 versus case A-2). As shown for the CVCV system in figure 7b, negligible additional savings (approximately 1 percent)

in chilled water energy consumption occur when an enthalpy economy cycle is

added to a zone-controlled cold deck (case B-2) as compared to one added to a fixed cold deck (case A-2). This result is largely due to the fact that a zone-controlled cold deck is effective during the winter months when solar

heat gains and outside-air induced heat conduction losses result in low cooling

loads. This condition may typically occur at outside air temperatures below

12.8°C (55°F) and it is therefore immaterial whether or not a zone-controlled

cold deck (setpoints ranging from 12.8°C (55°F) to 16.7°C (62°F)) is applied

since in this cold deck temperature range the cooling coil is nearly zero.

For the CV-CV system in figure 7b, reheat hot water energy savings from 12

percent to 21 percent are shown and are nearly identical to the savings shown

in figure 7a. Nearly equal savings are due to the fact that reheat hot water

energy consumption is dependent on the cold deck supply air set point temperature and not on the source for supply air cooling (outside air or cooling

coil)

26
For the VAV system, the results shown in figure 7b indicate additional savings

in chilled water energy consumption for the strategy consisting of zone- controlled cold deck with enthalpy economy cycle (case B-2) versus a fixed

cold deck with enthalpy economy cycle (case A-2). In contrast to the CV-CV

system, the VAV system can modulate the cooling capacity by resetting the zone supply air dampers. When all zones are experiencing moderate cooling loads the

supply air temperature for the CV-CV system is fixed at a certain value (ranging

from 12.8°C (55°F) to 16.7°C (62°F)) by the zone with the greatest cooling

demand. For the VAV system however, under similar zone loads, the supply air

temperature can be reset to a higher value since the zones air dampers will

increase the supply air quantity to satisfy the zone cooling loads. The net

result is that the VAV system operates at supply air temperatures which are more often higher than in the CV-CV system. By adding an enthalpy economy

cycle to the VAV system with zone-controlled cold deck (case B-2) the cooling

coil load is further reduced over a fixed cold deck with enthalpy economy cycle

(case A-2) since there are more hours in which the outdoor temperature is below

16.7°C (62°F) than below 12.8°C (55°F).

The savings in reheat hot water energy consumption for the VAV system are approximately equal to the results shown in figure 7a. The reason for the

nearly identical savings is the same as that given for the CV-CV system.

The results obtained for the VAV system and the CV-CV system in figure 7 were

found to apply as well to zone-controlled cold deck with temperature economy

cycle (cases B-2 versus A-3) in place of enthalpy economy cycle (case B-2 versus A-2).

6.1 EVALUATION OF SYSTEMS AND STRATEGIES

Analysis of the results obtained with the small office building have led to some general conclusions as to the performance of systems in small office

building of different thermophysical design and usage patterns. A more complete

discussion of practices and principles in HVAC system design and selection, may

be found in the ASHRAE Handbook and Product Directory^.

With respect to variable air volume systems, zones of a building which exhibit

small fluctuating loads may be well served by systems having a minimum air

volume ratio of 0.4 rather than 0.2. Presumably the former system is cheaper

on a first cost bases. Other considerations such as minimum ventilation

requirements and maximum outlet air temperatures may dictate the minimum air

volume ratio. Cold deck supply air temperatures must also be low enough to

permit adequate dehumidification of zones experiencing small sensible cooling

loads

In the following discussion some of the factors to be considered in efficient

zoning of two or more systems to a building are outlined. For optimum system

performance, a system should control those zones which exhibit similar heating

6 ASHRAE Handbook and Product Directory, 1980 Systems, American Society of

Heating, Refrigerating and Air Conditioning Engineers, New York, NY, 1980.

27
and cooling load patterns. This can permit effective use of cold deck supply

air reset and economy cycles.

When one or more zones must be continuously cooled it may be advisable to

dedicate one system to these zones. If these zones exhibit highly fluctuating

loads a variable air volume system or induction system may be appropriate. If

the loads are nearly constant during the occupied hours but are greatly reduced

during unoccupied hours, a constant volume system with intermittent operation

during unoccupied periods may be a feasable alternative.

In order to decide which system strategy is most appropriate it is necessary to

evaluate the relative energy operating costs of various combinations of systems

and strategies. This requires information of the cost of fossile fuels, electricity and the efficiencies of heating plants, chillers and circulating fans.

The cost of electricity and fossil fuels will vary by locality. In many cases

the type of heating plant, chiller and circulating fans is dictated by the system. This is evident, for instance for variable air volume systems in which

the supply fan should exhibit high efficiency at low supply air flow rates.

In this same system, the cooling coil may exhibit widely varying loads so that

the type of chiller and its part load performance can play a key role in

reducing energy costs.

28
N

3.7m

7. 3m

, 1.4m glass

rd Floor

2 . 3m

glass also placed on west-facing sides of zones 2 and 7

4 . 6m

1 and 2

n<^ Floors

A

Exterior Walls

Glass

Flat Roof

Occupants

Lights

Outlets

Infiltration

/

30.5m

: Metal Curtain/Airspace/2" Insul . /Gypsum Board

; Insulating; Sheet/Airspace/Sheet

: Built-up Roofing/2" Insul. /Metal Roof Deck

2 2

: 125 ft (11.6 m ) per person

: 2.5 W/ft 2 (26.9 W/m2

) ; 30% Return Air

: .3 W/ft 2 (3.2 W/m2

)

: 0 (Pressurized)

Figure 1 Small office building model

29
Hour

Electrical Appliances and Lighting Schedules

Figure 2 Internal heat gain schedules

30
HOT

WATER

CHILLED

WATER

ENERGY CONSUMPTION,

3

10 KWH

700

600

500

400

300

200

100

0

100

200

300 L

VAV System

CVCV System

CV System

nj 12

Figure 3 Annual heating and cooling energy consumption of three

HVAC systems utilizing the base case strategy (A-l) for

a small office building

31
HOT

WATER

CHILLED

WATER

ENERGY CONSUMPTION,

VAV System:

Strategy A-l

10

3

KWH E3 CVCV System: . Strategy B-2^ '

Figure 4 Comparative annual heating and cooling energy consumption

for a VAV system using the base case strategy with CV and

CV-CV system utilizing their respective best case strategies

( 1 ) Stragegy B-2 not possible to implement for Lake Charles and Santa

Maria for this system.

32
VAV System

0 CVCV System

PERCENT SAVINGS g CV System

CHILLED WATER

100 _

PERCENT INCREASE

HOT WATER

Figure 5 Percent savings in chilled water and percent increases in

hot water energy consumption for three HVAC systems using

enthalpy economy cycle*

*Average values for fixed cold deck temperature and zone controlled

cold deck temperatures

33

K\\\\\\\\\\\N\
PERCENT SAVINGS

VAV System

CHILLED WATER

100

80 "

60 -

40 -

{JyT| CVCV System

^

CV System

£ <u •u O t—

•H CO 4-1 CJ •rt 4-1 *0 CO

0) eu 0) to S ZC oP

80

100

PERCENT INCREASE,

HOT WATER

Figure 6 Percent savings in chilled water and percent increases in

hot water energy consumption for three HVAC systems using

temperature economy cycle*

*Average values for fixed cold deck temperature and zone-controlled

cold deck temperature.

34

Santa

Maria
PERCENT

SAVINGS

cno

5S<

m

H

!ZB

piw

fd

Figure 7a No Economy Cycle

Piw

<:QW

t-J i-4

U

Piw

<:

:sHO

60

40

20

20

40

60

Figure 7b Enthalpy Economy Cycle

Figure 7 Chilled water and hot water energy savings achieved for

strategies employing zone-controlled cold deck as compared

to a fixed cold deck temperature

35
Table 1. Climatological Data for Seven Cities*

City

Cooling

Degree Days

Heating

Degree Days

The percentage of temperature bin

hours ranging from representative

temperatures of 13.9°C (57°F) to

22.2°C (72°F) - for the daylight

hours from 9 a.m. to 4 p.m.

Lake Charles

,

Louisiana

2739 1498 31%

Nashville

,

Tennessee

1694 3696 34%

Washington, DC 1415 4211 32%

Dodge City,

Kansas

1411 5046 29%

Madison

,

Wisconsin

460 7730 28%

Seattle

,

Washington

183 4727 44%

Santa Maria,

California

84 3053 82%

.

* Derived from data supplied by: (1) Engineering Weather Data; Departments

of the Air Force, the Army and the Navy; AFM 88-29, July 1978; (2) Local

Climatological Data-Annual Summaries for 1975; National Oceanic and

Atmospheric Administration, Environmental Data Service, Ashville, N.C.

36
Table

2.

Comparative

Annual

Energy

Consumption

of

Three

HVAC

Systems

an

Selected

Strategies

fora

Small

Office

Building:

Lake

Charles 03

4-1 COSo

<Z)

<4H

NOm• i—

sf o•

f—H

CMO CO •

*-H -acm

O CO • • f-H fH<r o•

f-H

Of-H o•

f-H

'CI- ON. —H

-3-

1.015 |

41.9

1.015

f-H• CM

CMO•

f-H

f-H• CM

CMO•

f-H

90.6

2.191 o

i-J

344.4 o•

r-H 312.7 .908

341.0 .99

340.8 066*

308.6 968* 66*

o*ove

344.4

1.0

311.9

f-H ON•

496.8 |

1.443

<:> CO o O 00 00 o ON CO ON On CM r-4 o CO CO

m NO CM

CM • NO t-H CM • CM ON

m

O NO 00 • 00 • 00 • ,3 • f—H • o • rH • I"- • 00 • • NO • NO • O- ON i-T ON ON • ON in • m • m rH fH f-H • ^H f-H t—

f-H f-H rH f-H fH

m o o o o o o NO CO

CM • • • • • • C <3 <3 <3 < • r-. hJ o pH O *—

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ON ON On CM • s

CM CM 3

4-» to o f-H ON Ov O- 00

CO • • • o • On • o

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CC <3 CO CM

C/3

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CJ 00 o O ON CM On P" o

j= • • • <1* • CO < <! <3 < <3 • hJ 00 00 o NO O is is is is is *-h n*

00 ON • ON . fH •

f-H »-H f-H f-H f-H O' CO

ON o o' o O' o 00 NO o C"

O CO

m NO in NO CM MT

4h • • • • • • '3- CM NO CM • • CM • CM On •

f—H rH i-H ^H f-H f-H • r-' • ON • 00 !" 00 I-'. CM CM

CM CM CM 00 • 00 00 00 • oo •

f—H »—

f-H 00 00

rHS3 00 o UO o on n- CM 00 CM t-H f-H in <r 00 CM CM ON CM

4-4 CJ • • • pH • vO • CM • ON • 00 • o • ON f-H •

CO f-H CM 00 00 00 00 ON o n- 00 • NO • CO • CM

So <1- CM • in • ON • •-H • <1- On f-H On rH cn NO 440 m m m m NO in 1 o

CM ON o CO CM n- i-- <r nm CM co co ON CO VO o o • • • m • co • ON • <r • O • • • co CM • rC ON t-H ON o NO O ON ON o in • CO f-H • O' p-T n- 00 • 00 • • 00 • 00 *-H co •O fH

f—H »“H ^H f-H f-H f-H »—H rH f—H CM CM

CO CO

4-1 4-1 X) X)

CO 3 3

f—H no fcl 4-4 pH l 3 o c So p—l rH C CO r—H QJ E co CO co o o bO

-X r-'

O o 4-4 4-1 4-4 Pi Pi 3

(U CO Pi 3 4-4 4-4 tH

CO

ft! 4-4 (0 E o XJ X4 XJ So 3 3 3

cO Li co cx o CJ 0) 0) 04 e O

O

O

CJ s 04 f—H r-H rH

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a tH

X5 4-4 X) o t—i—

SO rH 3 1 | Sh 4-J a) 3 CO <0 CJ 04 o o E o o M Pi E •H

CO cO <0

X a) Pi Pi Pi o Pi CJ •H tH o XJ

cO

X tH 3 4-4 4J 3 4-1 0) 3 3 3 3 3 £ FO CO <0 4-1 Co 4-4 3 e O 3 O 3 O 3 u Pi cx c3 o o CJ O 04 Pi Pi o X CJ -X o x • *N r-H Pi CJ a 04 CJ Pi O O 3 4-J >> CO -u bo Pi CO 04 1 l 1 3 O O 3 CO bO o *H X3 tH J2 CX 04 04 SO 0) 4-4 x) 03 So co 3 3 r-H 0) c 04 CO 4-4 E 3 3 3- 3 3 AJ 4-4 3- O

4J > X s 04

OO rHO Pi 3 3 rHO 3

CO * *H

x tH Pi 04

U

N

N 3 N 3 OO 3 3 3

Pi 3 p> cj 4-4

O

X

a X. •H

4-1

O CO cO

O + + + + 4-4 + s + + 4-1 + PI CO •H r-l x II X) 3 3 3 3

4J (U 4-J 4-* f-H f-H f—H •-H 04 rH .U f—H f—H 3 f-H

O

a, pi a) co 3 l 1 i 1 1 1 1 1 CJ 0 i CO 4-4 O C C C + <3 + <3 <3 + < w 3 O

CO tH 3 4-1 f—H CM 1 co 1 f—H CM CO I f—H CM O CO | 1 1 | | 1 1 | O Pi c cl c PQ PQ PQ | o CJ a|

C

37

CV:

constant

volume

terminal

reheat

serving

all

zones

L^:

hot

water

CV-CV:

constant

volume

terminal

reheat

serving

different

zones Lc:

chilled

water

VAV:

variable-air

volume

with

terminal

reheat

serving

all

zones

Lf:

fan

(3)ts:

cold

deck

temperature
Table

3.

Comparative

Annual

Energy

Consumption

of

Three

HVAC

Systems

and

Selected

Strategies

fora

Small

Office

Building:

Nashville E3u

COto c

>

CJ1 uE

a) 4-> cotO cnl >

cj

tobo3

4-4 co3

•ucx

r-H o r-H o «—

o r-H o m o in OS

O Os in 00 o • O © o © co co CM 00 CNJ co in m Nf c CNJ in

4-1 • r-H 0 r-H • r-H 0 r-H e r-H

© f—

« o 0 0 9 oc

X sX vC vX CM 0 CM

© CM 0 00 0 00 i—

m 0

-d- <3- Nt

m f-H in r-H

m r-H r—H •Nf- 00 f—H

os

O o CO O' o CO 00 O' in c vO co vX in in vX o- • © « o © o. © ^d- o • f—H 0 in 0 0 00 o CNl f-H r-H CO co oc Os Os m O' o 00 r-H O' 0 in

X vX r-H

® CO 0 NT

© Os© Nt" 0 m 0 Os f—H 0 CNJ CNJ CM CM r-H CO CM f—H NT t-H O CNJ N* CO •-H in CNJ o Os Os O vX CO o « CO f-H o o CO Os in o <f O' m m 0 cn

X • r-H • o 0 o • vO 9 • vO 0 © 0 CM CM X o r-H o r-H 0 cx © 00 • 00 • O 0 o 0

*—

NT r-H CM CM CM CO CO ^H NT o C o o o O 00 o- 00 n- 00 n- 00 oc O' CM 4h • • • 9 • © • O' • Os © Os 0 Os 0 OS 9 CNJ X r—H f-H r-H r-H r-H r-H o Os o OS o Os o Os a os o o O o o « o 9 o 9 o 0 c 0 c 0

r-H r-H r-H r-H f-H f-H r-H CM CM m o sX CM co vC in co n- r-H Nfr m vX CM f-H CO e • • co e r-H 0 <1* o CM 0 c 0 m 0 CO e 00

CJ ex r-H 00 cx r-H Os vO vO nJ- C' vO os 00 vX Nf co

X OC o • "d- • vO 0 O • Nt-

© vO • o 0 vX 0 CO co Nt co co Nl" co r-H CM

r-H

r^. c CM vX CO CM NT r-H 00 o- CM <r O- Nt

C o X e e © vX • LO © 00 9 CO 0 CM 0 o 0 sX 0 c X co r-H O NT o O' oc o OS 00 Os Os vX Os CO in c r-H • r-H 0 i''. 9 Os 9 OC e 00 0 OS 9 r-H e CNJ CM r-H CM f-H r-H r-H r-H f—H r—

CO

CNJ o CM

O CM o CM o CM

O CM o CM o CM

C co vX <4H • • • • • e © CM 0 CM 0 eM 0 CM 0 CM Os

rJ O' r-H O' f-H Os r-H o O

O o O CM 0 CO C^ co o • o 0 o • o • o 9 f—H f"H r-H f-H f-H f-H f-H f-H f-H

CM o co t-H r-H Nf r-H co 'd* Os r-H CM r-H OS f—H NT CM OS

CJ • • © NT • CM 9 O' 9 co • CM

© CO 0 CO CO

hJ in 1—O vX co rvO 00 Os vX CM I-- in o CO CM 9 r—H co • o- 0 vX • CM • 0 co 0 00 ® f—H CM m co CO mT CO co m r—H co

CM LTl o CNJ NT VO r- CNJ CO 00 NT 00 r-' CNJ in 00

'W' • 0 « • 40 # NT 9 o • oc 0 vC 0 NJ- co r-' of

r—H CM o 00 o in r-H CM co f-H Os

m in n- 0

O'. f-H • o 0 CM 9 co

© CO • oo 0 o 9 co

r-H CNJ !~H CM r-H CM r-H CM f—

CM f—H CM f-H CO r-H m cn

4-1 4-1

33 33

03 3 3

t—H 33 U-4 4-J rH rH L 3 O 3 to T 1 rH c (U i—| 3 E CO CO cn O o bC

Hi 3 o 4-J 4-J 4-J 3 3 3

cu 3 3 >. 3 4-J 4-J •H

CO 3 4-J 3 E O 33

T3 33 3 3 3 3 & CtJ 3- O 3 3 3 3 E O O O

CJ 3 3 rH 1 1 iH o 3 3 •H

33 4-J 33

O i-l rH >s rH 3 1 1 4J

Oi 3 3 3 3 3 O o E o O 3 3 E •H

co 3 3 X 3 3 3 3 o 3 3 •H •H o 33 3 X •H 3 4-J 4-J 3 4-J 3 3 3 3 o 3

-3 CC CO 3 4-1 to 4-1 3 3 O 3 C 3 O 5 4-J 3 a 3 O O 3 O 3 3 3 3 XI 3

Hi o -3 • #\ rH 3 3 3 3 3 3 O

O 3 4-J CO 4-J bt • #s 3 3 3 1 1 I 3 O o 3 CO c •H nd •H X Cl 3 3 to 3 4-J 33 33 to W 3

f—H 0) 3 3 3 4-4 E 3 3 a 3 3 4-J 4-1 3. O > X 3 3 O O rH O 3 3 3 rH O 3

•H Hi •H 3 3 4-JNN 3 N 3 O

O cd 3 3 3 4-J 3 4-4 O X & X +

•H

O 3 3 O + + + + 4-J + E + + 4-1 3

•H rH

X ii 33 3 3 3 3 4-J 0) 4-J 4-J «—

f—f“H «—

3 f—

4-J r-H r-H 3 f—H

O 3- pii 3 CO 3 l l 1 l i 1 l 1 a E CO 4-J O < <3 <3 < + C + < < + < 'w' 3 o

co •rH 3 4-4 3

PC

«—CNJ CO r-H CM 1 CNJ 1 CJ <c 1 c 4c 1 PC 1 PC\ «l 1 Cj| 1 o| 4 o 3

aj

4-1 CO

)-i S 3

4-1 33

CO CU & iH

rH

4J *H o x x o X O

X

CO

cue

CO o

CU

N 3O X N rH

CO 4-t

CO

N CU >

>4-4 3

rH H-l 3

iH *H

co 33

bO bC co 3 3 0) -H *H X > > CU 3 3 3

cu a) CO CO

H

CO3

•HE

(-4 ai

4->

4-> 4-1 CO cC 0) 0) X X

CU 0) 3 3

rH iH

X

CO CO 4-> 3 3 -H •H *H 5

cu oi E

4-> 4-1 3

CU 0) E E 3 3

4-4 4J 3 C

co co

4-1 4-> co co C 3 O O 3 CJ

r> 3* 3* CP CJ < I > >

CJ

CO

38

cold

deck

temperature
Table

4.

Comparative

Annual

Energy

Consumption

of

Three

FVAC

Systems

and

Selected

Strategies

fora

Small

Office

Building:

Washington,

DC

cnE

a)

4-1 COto

COI >o o o o O o o CN rH o o 00 CN CO o NO ON 00 • • • • • • • CN • *H • rH • NO • o • NO

4-4 rH rH «-

H

rH rH »—H O' rH O' rH rH o • m • H m m m in • in • m • m • m r—1

00 rH

rH rH

CO o NO CO CO 00

m

O O' in NO NO 00 o 00 CN 00 • • • 00 • CN • CM • NO • • CO • NO • NO

CJ

m rH CNI 00 r^. ON i-'- NO in o O' O' • • H CO 1"- • ON • rH • m • r^- • CN • m ON rH

CNI *H r-H CN rH rH CN rH CO m o o rH o rH ON rH o co o CO NO co NO NO CO • • r-H rH *H r-H • co • co • CO m NO NO • X rH • o • o ON o- o o • O' • • CO • J in in • in • co • • <fr • rH • rH o CO tn rH m r-H CM o o m o O NO o O' O' NO o NO o r-H p'-

<4-4 • • • • • • • ON • O' • O' • ON • ON • r—

H

H cn •—1 co rH co rH co ON co O' CN O' CO ON CO ON NO 00

rH rH r-H rH • rH • rH • H • H e o • rH rH r-H rH rH rH rH rH CN rH

nO o in NO

m <r O' co o NO in rH r—H

NO CO n- • • • NO • • c • NO • co • rH • NO • p"

CJ rH rH o in p' NO ON ON 00 in co NO o O' rH in CO

hJ NO NO • ON • rH • in • O' • CN • NO • • CM CN CN CN CN rH CN

CN o NO 00

m co ON o o ON in rH

m CN

SP • • • NO • m • CO • O' • r^. • • O' • CO

t-J ON rH o r-H o CN 00 00 rH 00 CO 00m 00 co co

CM • • ON • c • o • O' • o • NO • CN CNI rH CM rH rH CN CN i—H

CM co

O o o o o o in

m

m

m in in

mm uo in 00 00

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cj • • • NO • • ON • 00 • NO • CN • O Pin rH 00 m 00 NO ro oc m 00 • rH ON NO NO rH • r-- NO • o • CN • • rH • 00 • rH CN

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33 Ph 4J 1 1 »H JL G o c to tH rH < QJ rH QJ E CO CO CO o o bO .X CJ o 4-J 4-J 4-1 p p 3

0) QJ >-i to C3 4-J 4-1 •H

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a

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c_>

C QJ rH »H rH o a CJ •H TO 4-J 03 o iH rH >- i—1 3 l 1 to 4-1 0) c co 0) CJ QJ o o E o O p P E •H

CO CO 0)

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pO

p CJ •H •HO 33

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O

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r—1

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a- c£ QJ CO

p 1 1 1 1 1 1 1 1 CJ E 1 CO 4-J o < <! C < + < + < <d + < P O

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p

a)

4-1 to

p S

a)

4-J 33

co ai

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C

O £ CO SP O <4-i X O 4-4 hJ J ,J

CO3g

CO o 3 N GO i—I N .-I

cO

n a)

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CO 33

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CO CO r-l

cO

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C

co CO *H

a> a) e

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qj ai ai

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CO CO 4-1 3 3 -H

•H *H ^ E E P P QJ 0)0)6

4J 4-1 3

r-H Q) QJ O B E > 3 3

»—1 i—I p O O -H > > coI 4-1 4-1 QJ G 3 r-l cO CO £2

4-1 4-1 (0 CO CO *H 3 3 J-l O O co o cj > > > > u o < I > >o

Q)PP

4-1 COP

QJaE

Q)

.x

cj

QJ 33

co

39
Table

5.

Comparative

Annual

Energy

Consumption

of

Three

HVAC

Systems

and

Selected

Strategies

fora

Small

Office

Building:

Dodge

City £

<D 4-> COtn

c/j %>

NO o m rm <r on 00 NO CO --H cn NO cn v£> m • CNJ CXn CM ON O' c 00 o —< o 0 O' • ON '<J' t-H

M-l • f-H • On • ON • o • o 0 0 0 <3- • • v*D H cO vD • vC • NO * NO 0 in in r-H • m in LP) in -h in ^H in O' —1

r—H o VO r-H no vc in NO NO

m cm On 00 no <r ON ON • • • m • 00 • n- • CO • oc 0 O' ® n- 0 cj On r-H IN- ® CM O' o CN ® CO ® r-H ® <3- NO

J in O' . t-H in • O'. 0 r-H in ON cn • CN r-H CN CM r-H CN CM r-H r-H

oc

O On CN CM ON 00

m cm

O in NO NO cn cn

r—H • f"- o 00 o cn c ON r—< O ON • ON • ON • ON

si 0 r-H • o & • • ON • O'. 0 ® ^H 0 t-H • J <3- • NT »—

00 • CO 0 O' NO NO CN • nO VO r-H vO m in in cm cn

NO c nD

O NOO cn oo co 00 cn oo cn oc cn oo r-H CM Si • • • • • • CTn • ON • ON 0 ON • a. 0 v£> H i—

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t—r-H »-H r-H t-H r—r-H r-H CM *-H

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C NO in v£>

® • • • • CM 9 • vO 9 CM

® • ON

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i—

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m o CM

M

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n

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in MD • VD r-H r-H • CM 0 CM r-H 0 CN • cn • CNl CN r-H CN CM CM CM CN CM oo cn

NT

O o- o o co n- NO O' oo cn in 00 <f 00 00 00

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® 0 CM • CM r-H • j nC r-H NO t—

V.0 r“

oc n- 00 00 • 00 r>- oo r". cn *-H n- CN • CN 0 CN CM ® CN •

rt—

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—i CM 00 00 cn n- in 00 00 On in

CJ • • • 00 • • ON 0 ON • NO 0 ON • NO o • hJ r—H r—t m m <J- nO m oo r-H in ON • o ON 00 NO cn cn

cm c VsO • r—

0 CM 0 * r-H LO CO CO CO cn in m

/-s CM <} o *—

00 ON

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C NO • • • • m • CM 0 00 • CM « r—

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cn

CO CD

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CO QJ 3

*—

xj Cm 4-) rH rH 1 c o 3 >N rH rH /-*“N < 0) i—H QJ E C/D CO CO O o bC o o 4-J 4-1 4-J S-l t-4 3

0) 3 M to 3 4-i 4-J •H

co 3 4-1 0) E O

X X?

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O

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C/D 3 3 X QJ

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0J c 04 3 4-i E 3 3 CL 3 3 4-J 4-1 CL o > X 3 0) O o rH O P 3 3 rH o 3 * *l“l •H H 0) 4-J N N cC N 3 o O 3 3 3 3 4-1 cj 4-1 O si CL , X •H O cu 3 o + + + + 4-i + E + -f- 4-J + ti °H t—H SI II X 3 QJ 3 3

4-1 QJ 4-J 4-J r-H r-H r-H r—

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O

a pi QJ co 3 1 I 1 1 1 L 1 l ° e i CO 4-1 O < < < < + < + < 4- <C W i 3 O

CO -H e 4-i r-H CN 1 7{ t—H

J CM CO I r—

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3

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4-> X! 3 3 & -H

i—

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s:

u

cn o 3 N 3O iH N rH3 m 4-1 3 3 bC 3 3 3 O t-i N 3 >

14-4 tJ

r-l l+J 3

rH •H cn 3 X 4-1 bC bC 3 3 3 3

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t-i t-4 tJ 3 3

cn cn rH3

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si J3

u 3 3 3

t-l t-i 4-1

r—i—

Si 3 3 4-1 3 3 •H

*H •H & E E

Lj t-i 3 3 3 E

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i—

t-i

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3 3 Si

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cn cn •H 3 3 t4 O O 3

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in3

4-> 3

t-i 0)CL£

01o

0) XX

rHO

CJ

si CJ 4-1 hJ hJ hJ

co

40
Table

6.

Comparative

Annual

Energy

Consumption

of

Three

HVAC

Systems

and

Selected

Strategies

fora

Small

Office

Building:

Madison

CO o co

O CO o O' O' in r- CO 00 in O' CM r—H CM 3 • <3 • 3 • o co o o co 3 CN c o O'

>4-1 • t-H • t-H • t-H • r—H • r-H • r—H • O • • • MT X co co CO O' • O' • O' • • i"- t-H o • m m in in r—H in r-H in r-H in i—H in 00 r-H £3

4-J CN o O' *<r co O' CM 3 CN CM r—H O' c r-H CN

CO • • • • • 00 • O' • O • 00 • 3 • o

to

O 00 r-H VC 3 <n 00 CM in CM O' 00 co • O' 00

ca

H oc CN • • 3 • r-H • co • 3 • r-H CO • ^H r-H rH r—H r-H r-H r-H i—H co t-H %>

co o r-H 3 r-H 3 00 O' O' r—H O' r-H 00 o O' CN 00 in

r-H • ro

O nO o O' r-H O' r—H CM O' 00 • 3

-3 • t—

• o • O • 00 • 00 • 00 • 00 • • in H VO • 3 • • • • 00 • CO 3 • O' O' r-H O' r-H r-^ r- CM CM

co o CO o CO o O 00 o 00

O 00 o 00

O 00 oc r—H

<4-4 • • • • • • • O' • O' • O' • O' • O' • 00

r-H r-H ^H i—^H r-H O' r—H O' r—H O' r-H O' ^H O' ^H in

CM CN CM CM • CM • CM • CM • CM • o. •

t-H •-H r-H r-H t—r-H r-H r-H t-H r—H E

4-4

C/D oc o o <fr CO

O O' o CM CO O' 3 O'

to • • • 00 • oc • 00 • 00 • 00 • • 00 • in

C/)

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00 CM 3

4-3 r\ 00 • CM • co • 00 • CM • co • OC • o • > co r—

CM CO t—

CM co r-H c CM

CJ' ^H

r—H o in 00 O' I-. r-H o CM 3 3 O' o X • • • <n • <r • • 00 • n- • CM • 3 • CO X O' r-H co o o o c 00 r—H 00 00 00 3 00 3 00 r—H r—H vC • 3 • r—

• 

CM • r-H • O • r—4 • oc • CM CM t-H CM t—

CM CM CM CM CM l-M co

CN o CN c CM o O'

m O' in 00 CO O' in O' in 00 co

>4-1 • CM • CM • r-. • CM • CM in in H 00 i—

00 r-H oc r-H r—H r-H n- CN • r-H r—H n- CN • 3 3 3 CM • CM • CM CM • CM • r—H

r-H r-H r-H r-H ^H r-H r-H i—H

r-H

"e3 o oc CM co

m CO to co <fr o co t-H O' 3 4-) o • • • O' • co • • o • 3 • CO • in 3 3 CO H CN i—

<* IT) <n O' 00 CM in r-H • in O' in

O • o O' • co • co • o • • CM • r-H CM

ca1 <r r-H CM CO CM CM co CM

CJ /*N CN LO o oc O' CO

n 00 3 o 3 CM r^. o o • • • 3 • in • r-H • • CO • in • r-H t—

m

,3 r-H ^H o o in CO 00 CO • <r r—H in 00 • X CO <r • • o • r—H • t—

r-H co • in • co

CN CM t—

CM r-H co r—H co r-H co co r-H co t—

co C/D 4-1 4J X X 3 3 3

r-H

X &h 4-J i-H t-H

L 3 o 3 to »-H rH /—V <C 3 ^H 3 E C/D co C/D o o bO

-X

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co £

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XX 3 3 3 3 3 a. o o 3 3 3 E O O O

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CO 3 3 X 3 P P P O

p o •H •HO -X

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O cj

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4-1 > X 3 3 O

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4-J

O 3 cdO + + + + 4-1 + E + + 4J + 4J

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X 3 3 3 3

4-1 3 4-J 4-J r-H r-H r—H r—H 3 r—H 4-J i—H r—H 3 i-H

O

a. cx 3 C/D 2 1 l 1 1 1 1 | i CJ E 1 c/d 4-J o < <3 <1 < + < + C < + <

C/D *H 3 4-J r-H CM CO r-H CM co 1 rH CM O 3 1 1 1 1 1 i i 1 CJ :X < < < ca CQ CQ | u o|

p

0)

4-1 CO P S

a)

4-1

X

co <u 5 i—

o X X CJ

CO33

COO

a)

n 3ON iH

4-1 3

0)P

beC

•P

n a; >

<4-4 P

iH <4-i a; i—I tH CO

co

x 4-1 be be co 3 3 3

tI *h

x > > 3 p p p 3 3

CO CO

4.) 4J

co <0 3 3 X X 3 0) P P

CO3

•HeP3

4-1 X

cO CO 4-1 3 3

•H i-l S E

(-< 3 QJ

4-> 4-1 3 <D E E 3 3

r—I i—

o o > >

4-1 4-1 3 3 3 co

X 4-J 4-J CO CO 3 3 O O

CJ

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CJ

01p3

4-1 3P

CDaE3

4-1

JxiCJ3X

,3 3 <4-1 X X X

co

41
Table

7.

Comparative

Annual

Energy

Consumption

of

Three

HVAC

Systems

Selected

Strategies

fora

Small

Office

Building:

Seattle

X3

CO

OC o 02 o m c CO CM in

O o r-H CO 20 UO r-H CM CM e 9 9 9 9 o r-H 02 r-H o r-H CM r-H r-H O 00

<4-4 • r—

H

r-H c- r-H 9 o 9 c 9 o 9 c 9 9 9 20

r*. CO 9 9 00 9 OC 9 00 ^H co 9 <T r-H t-H -'d' r-H r-H 00 r-H E LO o 02 r-H O CM 02 02 02 LO m r-H r-H 02 02 00 3 9 9 uo 9 r-H o r-H 9 9 OC in co 9 t—H u 3 CO r-H 9 <3* 20 uo 02 9 m 20 OC 9 9 CM 02 W l-J 20 9 r-H 9 m 9 00 9 o 9 *<! 9 CO 02 9 >2 r-H t—H

r-H 20 i—H

20 CO r—H CO>< m O <r 02 00 20 co 00 00 20 CO <fr o 00 20 > o 9 c r-H 9 r-H co r-H 00 02 00 m 20 CM 9 n-. 30 • t—H 9 o in o 9 00 9 00 9 00 9 r-» 9 9 OC CM CO <3- 9 m 0 20 9 r-- 9 20 9 O 9 r—H

CM 9 uo 02 r-H -d* r-H

CO o CO

O r-H uo <n 00 in r-2. UO I-. in 02 02

<4-4 • 9 9 9 9 o 9 02 9 02 9 02 9 02 9 02 9 r—H

l-J On ^H 02 r—

H

02 o 00 02 OC 02 00 02 00 02 00 02

G

O o 9 O 9 o 9 c 9 o 9 o 9 00 9

t-H r-H r-H r-H »-H 5—H

rH rH E

4-J 0) CM o <f 20 r-H CM 02

m CM 02 CO OC 02 <1- 20 00 • 9 02 CM 9 CO 9 m CM 9 9 02 CM 9 non o r—H 9 CM O CO CO 00 9 CM 20 co r-H 00 9 CM r-H 20 4 20 CM 9 9 r-H 9 9 co 9 »—H 9 CM 9 00 9 5

cn 00 «—

H co 00 rH co 00 02 CM

CV- u~) o 02

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m 20 in

m in 20 n- CM >d* 20 CM 9 9 9 o 9 00 9 00 9 20 9 9 9 uo 9

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r-H 02 9 co 9 20 9 OC 9 00 9 20 9 00 9 co co

CM CM r—H

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r-H 1—1

^H

CM o CM o CM c CO co co CO co CO, co CO CO co

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r-H r—H

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UO UO

m o 9 o 9 G 9 o 9 o 9 »-H *—* r-H r-H r-H r-H r-H r-H

"e3 o <T 20 20 in 20 i—H

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CA CM r—H

20 CM 02 20 00 20 CM 20 9 20 02 02 co c CM

02 02 e <n 9 r—

H 9 o 9 20 in 9 r-H 9 r-H 021 co r-H CO r-H r-H co r-H £

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m 20 00

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CO «—1

20 r-H r-H c o CM

O CO <3- CM 00 >d* r-H

m co

r-H co 9 co 9 20 9 00 9 © 02

© CM 9 CM CM r-H CM CM r-H CM CM r—H

CM r-H CO t-H W w

44 -U) X X

C/j 3 3

^H T3 En 4-J r-H rH

I, 3 o 3 r-H rH /—\ < 3 »-H 3 e w CO CD o o bO

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pp 3 3 3 P >2 3 4-J 44 •H 3 3 4-J 3 E O X X X 3 3 3 3 5 3 a. o 3 3 3 3 E O O O O 3 3 ' 1 i-H rH o 3 3 •H X 4-J X o t-H rH >2 rH 3 1 1 44 3 3 3 3 3 3 o o E o O P P E •H 3 3 3 X 3 P p p O p 3 •H •H o -X X 3 43 •H 3 4-J 4-4 3 3 3 3 3 3 3

43 CQ C/D 3 4-4 4-J 3 3 O 3 O cd

O s 4-J P 3- 3 O O 3 O 3 P P 3 40 3

44O

X • rH P 3 3 3 3 P O

O 3 44 CO 4-J bc • #v P cd 3 1 1 3 O O 3 C0 bC o •rl X •H 43 a 3 3 >2 3 4-J X X to W 3 3 r-H d) 3 3 3 4-> E 3 3 a 3 CO 4-J 4-1 a O

4-J > X 3 3 O

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Op 3 3 r-HO 3 3 * -H 4*2 •H P 3 4-4 N N 3 N 3 O o 3 3 3 P 3 4-1 3 4-1 O 43 P- 43 •H

4-J

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4-4 3 4-J 4-J *—

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<—I

r-H F—i 3 r-H 4-J r—

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r-H 3 r-H

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D- C*! 3 UD 3 i i 1 I 1 1 L 1 3 E 1 3 O 3 -H

C/D 4-J o < < < < + < + <c < + < 3 4-> rH CM CO r-H 1 CM 1 CO I r-H CM O 3 G Dd 1 < M i 05 1 03 1 U CJ Ql

p3

4-> cO P 5 3

4-4

X

CO 3

i—i

T—I

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43 3 <44 H H H

CM

W33

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CO 4J

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l—I *rl CO

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4-4 bC bC cO C C d) •H *H 43 > > 3 P P P

0) 3

CO CO I—I

cO

4-1 4-1 3

CO cO *H 3 <u e x x C

d) O) 3 P P 4-4

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X

CO CO 4-4 c 3 1-4 •H *H £ E P 3

d) d)

B

4-4 4-4 P

rH

Q) Q)O e b > 3 3 i—I —4 p O O t4 > > CO

4-1 4-4 (1) 3 3 x 3 3 40

4-4 4-13

CO CO "H 3 3 P 0 0 3 3 3 > > >, O < I > >o

CL' P3

4-4 3P3aE3

4-1 o

a> X

42
Table

8.

Comparative

Annual

Energy

Consumption

of

Three

HVAC

Systems

Selected

Strategies

fora

Small

Office

Building:

Santa

Mar

gg

GJCG to

GG

4-J tO

G 12 ai

4-1 CO <D

12 i—I

4-1 *H

C

O £ to j: a 144

43G 4-1

J tJ tJ

toa)c

CO o G N Co rH N rHG

co 4-J a> c be c G c o G •G N G >

'J-l

G

r-J <4-1

G

rH •H CO

cO

4-J b£ bo cfl C c G

•H •H 33 > > G G G G G G

CO CO 1—1 G

4-J 4-J 3

cfl cfl •H G G s

-C 33G

GGG

G

G 4-J { t““< j=

to cfl 4J c C •H

•H •H B E g H

G

GG £

4-> 4-) 3

rH

GG o £ B > 5 3

r—l

G o o •H > >

XJ 4-J

G c c r—J

cfl -O

4-J 4-J CO CO •H c c G o o G G G > > > > U u < 1 > >u

4-1 togg

p.BG

4-1 g

a) T3

T3

I—I cG

CO

43

APPENDIX A

RESULTS FROM SIMULATIONS OF ADDITIONAL STRATEGIES

Comparison of Additional Strategies for the CV System

Several additional strategies were simulated for the CV system in order to

determine the effects on energy consumption for hot water, chilled water and

fan energy.

Base Case Strategy (case A-l) and Base Case Strategy with Cold Deck Temperature

of 55°F (12 ,8°C) (cases F-l , F-2)

The results were obtained for Washington, DC weather only to illustrate the

effects of reducing the cold deck temperature from 16.7°C (62°F) to 12.8°C

(55°F) (cases A-l and F-l respectively). Strategy F-2 is similar to strategy

F-l except that an enthaply economy cycle is added.

For strategy F-l the hot water energy consumption increases by 12 percent and

chilled water consumption increases by 4 percent over the base case. Due to

lower air flow rates required at a 12.8°C (55°F) supply air temperature, a reduction in fan energy of approximately 31 percent is observed.

Incorporating an enthalpy economy cycle to strategy F-l (case F-2) results in an increase in hot water energy consumption of 19 percent over the base case,

and identical fan energy consumption as in case F-l . Reduction in chilled

water energy consumption over the base case is 34 percent. It should be noted,

however, that larger savings are obtained by employing a 16.7°C (62°F) cold

deck with enthalpy economy cycle (case B-2) and amounts to 44 percent.

Effect of Reducing the Cold Deck Throttling Range from 1.7°C (3°F) to 0.28°C

(0 .5°F) (case H-l

)

Strategy H-l was similar to the base case strategy with an enthalpy economy

cycle added (case A-2) except that 0.28°C (0.5°F) cold deck throttling range

was simulated. This case was simulated for Washington, DC weather. Although

such a small throttling range may be impractical, the results are used to

illustrate the sensitivity of hot water and chilled water energy consumption to

throttling range. As compared to strategy A-2 hot water and chilled water

energy consumption decreases by 29 percent and 14 percent respectively - fan

energy consumption remains the same.

Comparison of the Base Case Strategy to One Employing a Cold Deck Temperature

of 12.8°C (55°F) and Reheating at 25.6°C (78°F) (case E-l)

It is evident strategy E-l is extremely wasteful, the extent in increases in

energy consumption over the base case are listed below.

Fan Energy Consumption : Due to reduced air flow rates required when operating

with a cold deck temperature of 12.8°C (55°F) decreases in fan energy over the

base case range from 30 percent to 33 percent for all seven cities.

A-l
Hot Water Energy Consumption : Increases In hot water energy consumption over

the base case strategy range from 60 percent to 80 percent, decreasing with a decrease in the number of cooling degree days for all seven cities.

Chilled Water Energy Consumption : Increases in chilled water consumption over

the base case range from 15 percent to 21 percent for all seven cities.

Comparison of a VAV System with a Minimum Air Volume Ratio of 0.4 (cases G-l

and C-2) with the Base Case Strategy ( A— 1

)

Simulations were conducted for the VAV system for a minimum air fraction of 0.4

with an enthalpy economy cycle (case G-2) and without an enthalpy economy cycle

(case G-l). In all other respects the strategies for this system are the same

as the base case strategy (case A-l). Energy consumption estimates were made

for several cities based on correlation obtained from the results of the

remaining cities and are presented in table A-l.

Fan Energy Consumption . Increases in fan energy consumption for strategies G-l

(without enthalpy economy) and G-2 (with enthalpy economy) over the base case strategy (case A-l) range from 8 percent to 10 percent for six cities, and

approximately 13 percent for Dodge City. As observed in previous results

negligible differences in fan energy consumption result when adding enthalpy

economy cycles (case G-2).

Hot Water Energy Consumption . The following table lists the percent increases

in hot water energy consumption required for reheat for strategy G-l compared

to the base case strategy (A-l). The second row compares the base case strategy

A-l for the VAV system with the corresponding base case strategy (A-l) for the

CV system.

Percent

Increase

Lake

Charles Nashville

Washington

DC Dodge Madison Seattle

Santa

Maria

(Gl-Al ——— x 100 154% 80% 65% 64% 32% 70% 158%

A1 VAV

( A1 r>VT~A1

y

ATT ) GV VAV 100 833% 384% 288% 288% 141% 302% 1082%

>%

I c—H

The above table indicates that the increases in hot water energy consumption of

strategy G-l with respect to A1 for the VAV system follow the same trend with

cities as the increases in base strategy A-l for the CV system with respect to

the base case in the CV system. This is expected since as the extent of maximum damper air closure decreases, the system approaches a constant volume

system.

Adding an enthalpy economy cycle to strategy G-l (case G-2) results in a 3 to 8

percent increase in hot water energy consumption. The percent increases in hot

water energy consumption generally decrease with a decrease in the number of

Acooling degree days except for Seattle and Santa Maria. It is evident that

other factors influence the increase in hot water. As discussed previously,

these factors can include the daily outdoor temperatures extremes as well as the mean daily temperature for each season of the year, the length of seasons

and the amount of solar radiation.

Chilled Water Energy Consumption . The chilled water energy consumption

increases for strategy G-l (without enthalpy economy cycle) and strategy G-2

(with enthalpy economy cycle) with respect to case A-l are listed below for

the seven cities.

Percent Lake Washington Santa

Increase Charles Nashville DC Dodge Madison Seattle Maria

Case G-l 18% 17% 20% 19% 22% 32% 16%

Case G-2 4% -12% -19% -17% -20% -50% -37%

The percent increases in strategy (G-l) (without enthalpy economy cycle) as compared to strategy A-l range from 16 percent to 32 percent for the seven

cities. For case G-2, Lake Charles still indicates higher chilled water energy

consumption than base case A-l (without enthalpy economy). It is evident for

this city, outdoor temperatures are not low enough or prolonged to result in

decreases in chilled water consumption. The remaining cities, however, indicate significant decreases in chilled water consumption, the largest increases

occurring for Seattle and Santa Maria. The primary reason for the large

decreases exhibited in these cities has been discussed previously. It should

be noted, however, that chilled water savings are greater for all cities for

strategy A-2 (enthalpy economy and minimum air volume ratio of 0.2) than for

strategy G-2.

ATable

A-l

.

Comparison

of

Energy

Consumption

for

VAV

Systems

with

Minimum

Air

Volume

Ratios

of

0.2

and

0.4

for

Seven

Cities

A-4
NBS-114A (REV. 2 - 80 ) U.S. DEPT. OF COMM.

BIBLIOGRAPHIC DATA

SHEET (See instructions)

1. PUBLICATION OR

REPORT NO.

NBSIR 82-2489

2. Performing Organ. Report No. 3. Publication Date July 1982

4. TITLE AND SUBTITLE

Strategies for Energy Conservation in Small Office Buildings

5. AUTHOR(S)

Walter H. Parken, James Y. Kao, George E. Kelly

6. PERFORMING ORGANIZATION (If joint or other than NBS. see in struction s)

national bureau of standards

DEPARTMENT OF COMMERCE

WASHINGTON, D.C. 20234

7. Contract/Grant No.

8. Type of Report & Period Covered

9. SPONSORING ORGANIZATION NAME AND COMPLETE ADDRESS (Street. City. State. ZIP)

Office of Buildings Energy Research and Development

Office of Conservation and Renewable Energy

U.S. Department of Energy

Washington, D.C. 20585

10. SUPPLEMENTARY NOTES

~| Document describes a computer program; SF-185, FIPS Software Summary, is attached.

11. ABSTRACT (A 200-word or less factual summary of most significant information. If document includes a significant

bibliography or literature survey, mention it here)

A comparative analysis is made of the thermal performance of a small office building

using various HVAC systems and commonly employed strategies. The comparisons are made

for seven geographical locations representing wide climatic variations within the conti

nental United States.

Results were obtained for fan, space heating hot water, and chilled water energy con- sumption through hour-by-hour simulations using the BLAST computer program. A small

office building model was used in the simulations along with several HVAC systems; a constant volume reheat unit (serving the entire building), dual constant volume reheat

units (serving separate zones of the building), and a variable air volume reheat unit.

The strategies investigated included supply air temperature reset (constant, zonecontrolled, and outdoor air-controlled), economy cycles (temperature and enthalpy),

continuous conditioning versus conditioning only during occupied hours, changes in re- heat set point temperature, and changes in minimum variable air volume ratio.

For comparable control strategies, the variable-air volume terminal reheat system exhibited the least energy consumption for chilled water, hot water and circulating fan.

The system incorporating two independently-operating constant volume terminal reheat

units ranked second in energy consumption while the single constant volume terminal reheat unit ranked last. Changes in thermal performance resulting from implementing one

strategy in place of or in combination with another were found to vary significantly by

climate and the type of HVAC system employed.

12. KEY WORDS (Six to twelve entries ; alphabetical order; capitalize only proper names; and separate key words by semicolons)

Building control strategies; building energy conservation; building thermal performance

HVAC

13. AVAILABILITY

jXX| Unlimited

| | For Official Distribution. Do Not Release to NTIS

Order From Superintendent of Documents, U.S. Government Printing Office, Washington, D.C.

20402.

[KXj Order From National Technical Information Service (NTIS), Springfield, VA. 22161

14. NO. OF

PRINTED PAGES

55

15. Price

$9.00

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