NEW NIST PUBLICATION

February 7, 1989

NISTIR 88-3860

The Interaction of Lighting,

Heating and Cooling

Systems in Buildings - Interim Report

Stephen J. Treado and John W. Bean

U.S. DEPARTMENT OF COMMERCE

National Institute of Standards and Technology

(Formerly National Bureau of Standards)

Center for Building Technology

Building Environment Division

Gaithersburg, MD 20899

September 1988

Issued December 1988

Prepared for

U.S. Department of Energy

Office of Buildings and Community Systems

NISTIR 88-3860

The Interaction of Lighting,

Heating and Cooling

Systems in Buildings - Interim Report

Stephen J. Treado and John W. Bean

U.S. DEPARTMENT OF COMMERCE

National Institute of Standards and Technology

(Formerly National Bureau of Standards)

Center for Building Technology

Building Environment Division

Gaithersburg, MD 20899

September 1988

Issued December 1988

National Bureau of Standards became the

National Institute of Standards and Technology

on August 23, 1988, when the Omnibus Trade and

Competitiveness Act was signed. NIST retains

all NBS functions. Its new programs will encourage

improved use of technology by U.S. industry.

U.S. Department of Energy

Office of Buildings and Community Systems

^w A

^ris o*

Prepared for

U.S. DEPARTMENT OF COMMERCE

C. William Verity, Secretary

NATIONAL INSTITUTE OF STANDARDS

AND TECHNOLOGY

Ernest Ambler, Director

Abstract

The effect of interactions between building lighting, heating and

cooling systems on the energy performance of the lighting system and

cooling loads is examined, based on detailed full-scale measurement and

supporting computer simulations. A test facility was designed,

constructed and operated to emulate an office space with recessed

fluorescent lighting. Special design features simulated adjacent

building areas, providing realistic thermal boundary conditions for the

test room. The test facility was extensively instrumented to monitor

lighting power, cooling load, surface and air temperatures, heat flows

and light levels. 398 measured parameters were averaged and recorded

every two minutes during testing.

This interim report describes preliminary results from the research

effort. The results showed that the lighting system can be constrained

to operate at its most efficient level, if the fluorescent lamps are

cooled sufficiently.

A two-term exponential relation with four regression coefficients was

found to fit the measured data well. An extended set of weighting

factors was derived based on the double exponential relation, and values

computed from the measurements. Cooling load profiles due to cyclic

operation of the lights were computed using the weighting factors.

Lighting energy distribution fractions were determined using a

calibrated computer model. Design considerations are discussed.

Keywords

building energy, cooling, heat transfer, lighting, peak

loads, weighting factors
Acknowledgement

The support of the U.S. Department of Energy and the Electric Power

Research of Energy and the Electric Power Research Institute to this

research effort is greatfully acknowledged. The useful discussions with

Reinhard Radermacher, and Tamami Kusuda are appreciated. The

contribution of Ross and Baruzzini and Dubin-Bloome Association in

drawing the test facility plans is acknowledged. Thanks to Beth

Burdette for typing and editing the manuscript.

Disclamer

Certain manufacturers and trade names are mentioned in this report for

the purpose of describing the testing facility and equipment. This does

not constitute an endorsement or recommendation of such equipment, but

is provided for informational purposes only.

11
Table of Contents

Page

List of Tables iv

List of Figures v

Nomenclature. xii

I . Introduction. 1

2 . Background 2

2.1 The Temperature Dependence of Fluorescent Lamps ....2

2 .

Energy Distribution from Fluorescent Luminaires ......... 3

2 .

Cooling Load Due to Lighting. 5

2.4 Weighting Factors and Design Procedures. .6

2. 5 Building Energy Analysis Computer Programs .............. 8

2 .

Summary of Previous Work 10

2.7 Objective of the Present Research.. 12

3 . Experimental Facility and Plan. 18

3.1 Test Facility .....18

3.2 Instrumentation 19

3 . 3 Measurements and Calibration. 22

3.4 Test Parameters 24

4. Computer Model of Lighting and HVAC Interaction 38

4. 1 Basis of Computer Model. 38

4 . 2 Solution Procedure 39

5. Steady State Measurement Results 43

5.1 Lamp Compartment Return 43

5.2 Effect of Return Configuration 46

5 . 3 Luminaire Temperatures 47

6. Transient Measurement Results 61

6.1 Cooling Load Profiles and Regression Analysis 61

6.2 Weighting Factors and Cooling Load Fractions 67

6 . 3 Heat Storage in Room Components 70

6.4 Plenum Cooling Load Fraction .72

7. Lighting Energy Distribution Fractions 113

7.1 Calibration and Validation of Computer Model 113

7.2 Computer Simulation Results 114

8. Discussion of Results 121

9. Conclusions 123

10. Appendix 125

II. References.... 132

iii
List of Tables

Page

Table 2.1 Energy Distribution Fractions From Lighting . Systems ..

Table 3.1 Test Parameters 24

Table 4.1 Nodes for Test Facility Simulation 38

Table 4.2 Convection Correlations for Simulations... .....40

Table 6.1 Regression Coefficients for Sideslot, 200cfm Tests.. 65

Table 6.2 Average Regression Coefficients for each Test

Condition. 66

Table 6.3 Average Weighting Factors for each Test Condition. .. 67

Table 6.4 Lighting Power and Peak Cooling Load. 68

Table 6 . 5 ASHRAE-Type Weighting Factors 69

Table 6.6 Heat Storage Fractions in Room Components .71

Table 7.1 Lighting Energy Distribution Fractions .115

iv
List of Figures

Page

Figure 2.1 Fluorescent Lamp Light Output and Power 14

Consumption Versus Minimum Lamp Wall Temperature

Figure 2.2 Schematic Diagram of Fluorescent Lamp Components .... 15

Figure 2.3 Energy Distribution Fractions from Lighting Systems. 16 .

Figure 2.4 Cooling Load due to Step Change in Lighting. 17

Figure 2.5 Cooling Load with Cyclic Operation of Lighting 17

System

Figure 3.1 Cut-away Schematic View of Test Facility 25

Figure 3.2 Floor and Ceiling Slabs are Supported by Steel 26

Frame

Figure 3.3 Guard Air Spaces Surrounding Test Room 27

Figure 3.4 Lighting Systems... 28

Figure 3.5 Access Doors Allow Entry to Guard Air Spaces 29

Figure 3.6 Test Room Air Supply and Return Ducts, and Guard.... 30

Air Ducts

Figure 3.7 Duct Layout in Lower Plenum 31

Figure 3.8 Test Room Air Temperature Thermocouple Grid 32

Figure 3.9 Surface Temperature Measurement Grid 33

Figure 3.10 Heat Flux Transducer Locations 34

Figure 3.11 Wall Guard Air Thermopile Control System Schematic .. 35

Figure 3.12 Cooling Load Oscillations Due to Excessively High... 36

Control Loop Gain Constants

Figure 3.13 Heater Control Signal Oscillations Due to 36

Excessively High Control Loop Gain Constants

Figure 3.14 Outputs From Supply - Return Air Thermopile and.... 37

Wall Thermopile During Transient Test

Figure 3.15 Wall Thermopile Outputs on Expanded Scale 37

Figure 5.1 Lighting Power Versus Room Air Temperature and 48

Airflow Rate for Lamp Compartment Return

Figure 5.2 Lighting Power Versus Average Minimum Lamp Wall 48

Temperature

v
Figure 5 . 3 Light Output Versus Room Air Temperature and . . . 49

Airflow Rate

Figure 5.4 Light Output Versus Average Minimum Lamp Wall. ..... .49

Temperature

Figure 5.5 Lighting System Luminous Efficacy Versus Average .... 50

Minimum Lamp Wall Temperature

Figure 5.6 Average Minimum Lamp Wall Temperature Versus Room... 50

Air Temperature and Airflow Rate

Figure 5.7 Average Plenum Air Temperature with Lights on....... 51

Versus Room Air Temperature and Airflow Rate

Figure 5.8 Average Floor Temperature with Lights on Versus ..... 51

Room Air Temperature and Airflow Rate

Figure 5.9 Average Ceiling Temperature with Lights on Versus... 52

Room Air Temperature and Airflow Rate

Figure 5.10 Average Wall Temperature with Lights on Versus ...... 52

Room Air Temperature and Airflow Rate

Figure 5.11 Average North Wall Temperature with Lights on....... 53

Versus Room Air Temperature and Airflow Rate

Figure 5.12 Average East Wall Temperature with Lights on........ 53

Versus Room Air Temperature and Airflow Rate

Figure 5.13 Average South Wall Temperature with Lights on 54

Versus Room Air Temperature and Airflow Rate

Figure 5.14 Average West Wall Temperature with Lights on........ 54

Versus Room Air Temperature and Airflow Rate

Figure 5.15 Lighting Power Versus Airflow Rate and Path. 55

Figure 5.16 Average Minimum Lamp Temperature Versus Airflow 56

Rate and Path

Figure 5.17 Plenum Air Temperature Versus Airflow Rate and 56

Path

Figure 5.18 Ceiling Temperatures Versus Airflow Rate and Path... 57

Figure 5.19 Floor Temperatures Versus Airflow Rate and Path..... 57

Figure 5.20 North Wall Temperature Versus Airflow Rate and...... 58

Path

Figure 5.21 East Wall Temperature Versus Airflow Rate and Path.. 58

vi
Figure 5.22 South Wall Temperatures Versus Airflow Rate and 59

Path

Figure 5.23 West Wall Temperatures Versus Airflow Rate and 59

Path

Figure 5.24 Luminaire Temperatures During Lights on Step ........ 60

Change Test

Figure 5.25 Luminaire Temperatures During Lights off Step ....... 60

Change Test

Figure 6.1 Measured Cooling Load and Lighting Power for .73

Back- to -Back Lights On and Lights Off Tests

Figure 6 . 2 Adjusted Cooling Loads 73

Figure 6.3 Measured Cooling Load and Lighting Power Versus ..... 74

Single Exponential Regression Fit, CG160 On

Figure 6.4 Calculated Minus Measured Cooling Load, CG160 On.... 74

Figure 6.5 Measured Cooling Load Versus Double Exponential 75

Regression Fit, CG160 On

Figure 6.6 Measured Cooling Load and Lighting Power Versus 75

Single Exponential Regression Fit, CG160 Off

Figure 6.7 Initial Transient, CG160 Off.. ...76

Figure 6.8 Measured Cooling Load Versus Double Exponential ..... 76

Regression Fit, CG160 Off

Figure 6.9 Measured Cooling Load and Lighting Power Versus 77

Single Exponential Regression Fit, CG120 On

Figure 6.10 Measured Cooling Load and Lighting Power Versus 77

Double Exponential Regression Fit, CG120 On

Figure 6.11 Measured Cooling Load and Lighting Power Versus 78

Single Exponential Regression Fit, CG200

Figure 6.12 Initial Transient, CG200 78

Figure 6.13 Measured Cooling Load Versus Double Exponential 79

Regression Fit, CG200

Figure 6.14 Measured Cooling Load and Lighting Power Versus 79

Single Exponential Regression Fit, SS120 On

Figure 6.15 Initial Transient SS120 On... 80

Figure 6.16 Measured Cooling Load Versus Double Exponential 80

Regression Fit, SS120 On

vii
Figure 6.17 Measured Cooling Load and Lighting Power Versus 81

Single Exponential Regression Fit, SS120 Off

Figure 6.18 Initial Transient, SS120 Off. .81

Figure 6.19 Measured Cooling Load Versus Double Exponential ..... 82

Regression Fit, SS120 Off

Figure 6.20 Measured Cooling Load and Lighting Power Versus 82

Single Exponential Regression Fit, SS160 On

Figure 6.21 Initial Transient, SS160 On. 83

Figure 6.22 Measured Cooling Load Versus Double Exponential ..... 83

Regression Fit, SS160 On

Figure 6.23 Measured Cooling Load and Lighting Power Versus ..... 84

Single Exponential Regression Fit, SS160 Off

Figure 6.24 Initial Transient, SS160 Off. . 84

Figure 6.25 Measured Cooling Load Versus Double Exponential ..... 85

Regression Fit, SS160 Off

Figure 6.26 Measured Cooling Load and Lighting Power Versus ..... 85

Single Exponential Regression Fit, SS200 On

Figure 6.27 Initial Transient, SS200 On. .......... 86

Figure 6.28 Measured Cooling Load Versus Double Exponential 86

Regression Fit, SS200 On

Figure 6.29 Measured Cooling Load and Lighting Power Versus 87

Single Exponential Regression Fit, SS200 Off

Figure 6.30 Initial Transient, SS200 Off ...87

Figure 6.31 Measured Cooling Load Versus Double Exponential 88

Regression Fit, SS200 Off

Figure 6.32 Measured Cooling Load and Lighting Power Versus 88

Single Exponential Regression Fit, LC120 Off

Figure 6.33 Initial Transient, LC120 Off............ ..89

Figure 6.34 Measured Cooling Load Versus Double Exponential 89

Regression Fit, LC120 Off

Figure 6,35 Measured Cooling Load and Lighting Power Versus 90

Single Exponential Regression Fit, LC160 On

Figure 6.36 Initial Transient, LC160 On. .90

Figure 6.37 Measured Cooling Load Versus Double Exponential 91

Regression Fit, LC160 On

viii
Figure 6.38 Measured Cooling Load and Lighting Power Versus 91

Single Exponential Regression Fit, LC160 Off

Figure 6.39 Initial Transient, LC160 Off 92

Figure 6.40 Measured Cooling Load Versus Double Exponential 92

Regression Fit, LC160 Off

Figure 6.41 Measured Cooling Load and Lighting Power Versus 93

Single Exponential Regression Fit, LC200 Off

Figure 6.42 Initial Transient, LC200 Off 93

Figure 6.43 Measured Cooling Load Versus Double Exponential 94

Regression Fit, LC200 Off

Figure 6.44 Comparison of Cooling Load Profiles From 94

Different Tests with Same Conditions

Figure 6.45 Regression Coefficients Versus Airflow Rate 95

Figure 6.46 Regression Coefficients A, B and C Versus Airflow... 95

Figure 6.47 Regression Coefficient A Versus Airflow 96

Figure' 6.48 Regression Coefficient B Versus Airflow 96

Figure 6.49 Regression Coefficient C Versus Airflow 97

Figure 6.50 Regression Coefficient D Versus Airflow 97

Figure 6.51 Cooling Load Profiles for Ceiling Grill Return 98

Figure 6.52 Cooling Load Profiles for Ceiling Grill Return 98

Figure 6.53 Cooling Load Profiles for Ceiling Lamp 99

Compartment Return

Figure 6.54 Cooling Load Profiles for Ceiling 120 cfm Airflow... 99

Figure 6.55 Cooling Load Profiles for Ceiling 160 cfm Airflow.. 100

Figure 6.56 Cooling Load Profiles for Ceiling 200 cfm Airflow.. 100

Figure 6.57 Cooling Load One Hour After Lights are Energized. .. 101

Figure 6.58 Weighting Factor a0 , Versus Airflow 101

Figure 6.59 Weighting Factor a^ Versus Airflow 102

Figure 6.60 Weighting Factor a.2 Versus Airflow 102

Figure 6.61 Weighting Factor b]_ Versus Airflow 103

IX
Figure 6.62 103

Figure 6.63

Figure 6.64

Figure 6.65

Figure 6,66

Figure 6.67

Figure 6.63

Figure 6.69

Figure 6.70

Figure 6.71

Figure 6,72

Figure 6.73

Figure 6,74

Figure 6,75

F i gur e 6,76

Figure 6.77

Figure 6.78

Figure 6,79

Figure 7,1

Weighting Factor Versus Airflow

Hourly Cooling Load Fractions for Four Days 104

Hourly Cooling Load Fractions Versus Airflow Rate.. 104

for Ceiling Grill

Hourly Cooling Load Fractions Versus Airflow Rate.. 105

for Side Slots

Hourly Cooling Load Fractions Versus Airflow Rate.. 105

for Lamp Compartment

Hourly Cooling Load Fractions Versus Airflow Rate.. 106

for Average Conditions

Cooling Load Profiles Versus Airflow for Ceiling ... 106

Grill

Cooling Load Profiles Versus Airflow for Side 107

Slots

Cooling Load Profiles Versus Airflow for Lamp ...... 107

Compartment

Cooling Load Fractions with Original and Adjusted. . 108

Weighting Factors, CG200

Cooling Load Fractions with Original and Adjusted, . 108

Weighting Factors, SS160

Cooling Load Fractions with Original and Adjusted, . 109

Weighting Factors, SS160

Total Heat Storage Versus Airflow Rate and 109

Return Configuration

Heat Flow Sensor Readings for Floor and Walls 110

During Back-to-Back Lights on and Lights off Tests

Plenum Heat Pickup Fraction Versus Airflow Rate.... 110

and Return Air Configuration

Plenum Fraction Versus Time for 160 cfm, Ceiling. .. Ill

Grill Return

Plenum Fraction- Versus Time for 160 cfm, Side Ill

Slot Return

Plenum Fraction Versus Time for 160 cfm, Lamp. 112

Compartment Return

Comparison of Measured and Simulated Cooling 116

Loads, Ceiling Return, 120 cfm

x
Figure 7.2 Comparison of Measured and Simulated Cooling 116

Loads, Ceiling Return, 200 cfm

Figure 7.3 Comparison of Measured and Simulated Cooling.. 117

Loads, Lamp Compartment Return, 120 cfm

Figure 7.4 Comparison of Measured and Simulated Cooling ..117

Loads , Lamp Compartment Return 200 cfm

Figure 7.5 Lighting Energy Distribution Fractions Versus 118

Time, Ceiling Return, 120 cfm

Figure 7.6 Lighting Energy Distribution Fractions Versus 118

Time, Lamp Compartment Return, 120 cfm

Figure 7.7 Lighting Energy Distribution Fractions Versus ...... 119

Time, Ceiling Return, 120 cfm, 80°F

Figure 7.8 Cooling Load Profile for Heavy Thermal Mass 119

Room, Ceiling Return, 120 cfm

Figure 7.9 Lighting Energy Distribution Fractions Versus ...... 120

Time, Heavy Thermal Mass Room, Ceiling Return,

120 cfm

xi
Nomenclature

A, B , C

,

D - curve fitting constants

AF - supply and return airflow rate

A^ - area of node i

aQ, a^ , a2 ,b^ ,b2 - weighting factors

- radiosity of node i

C,C0

- tracer gas concentrations

- thermal capacitance at node i

CK^ - conduction heat gain at node i

Cp - specific heat of air

e - error signal

EB^ * black body equivalent thermal radiation power of node i

EM-j_ - emittance of node i

Fj_j * radiation shape factor between nodes i and j

G - transfer function

h - convection coefficient

HC^ - convection coefficient at node i

Hi - irradiance at node i

I - infiltration rate

k - sample number

Kj - integral gain constant

K-jj - conductance between nodes i and j

Kp - proportional gain constant

m - mass

m - mass flow rate of supply air

Q

- cooling load

QC^ - convection heat gain at node i

xii
QRN^ - net thermal radiation at node i

QS^ - heat storage at node i

QVi - visible light watts incident on node i

s - Laplace transform variable

Sj. - fraction of visible light absorbed by node i

t - time

T - time increment

Tair " air temperature

TAO - time increment

Ti - temperature of node i

TjiLW " minimum lamp wall temperature

Tret " return air temperature

Tsup

“ supply air temperature

VW - light output in watts

W - lighting power

z - z transform variable

Am - change in control signal

AT - temperature difference

a - Stefan - Boltzmann constant

p - density of standard air

xiii
-
I . Introduction

Lighting in commercial buildings is the single largest user of electric

energy, typically ranging from 25 to 50 percent of total building

electrical energy requirements. In 1980, 2.8 ( 10 1

) Btu of energy was

used for lighting for a total cost of over 40 billion dollars

annually [Un85 ] . An overall increase in lighting efficiency of one

percent would produce a savings of 400 million dollars a year.

The performance of the dominant commercial light source, the fluorescent

lamp, is strongly dependent on thermal conditions, with both lamp light

output and power consumption varying with minimum lamp wall temperature

as much as 20 percent under typically encountered conditions. Proper

control of room thermal conditions can ensure that the lamps are

operating at their most efficient level.

In addition to the electrical energy purchased and used for lighting

itself, heat dissipated from the lighting system adds to the building

cooling load in summer, and decreases the heating requirements in

winter. Controlling peak cooling loads in summer is of particular

interest to electric utilities which have to reduce summer peaks at any

cost to control peak electric power demands on hot summer afternoons.

Due to finite electric power generation resources, steady electric power

demands make most efficient use of electric power generation facilities.

The capital cost of expansion of generating capacity has, in turn, led

to greater costs to the user in the form of demand charges and ratchet

clauses. A demand charge usually takes the form of a higher unit cost

for electrical power during periods of heavy system - wide usage.

Demand charges frequently are assessed from late morning through early

evening

A ratchet clause ties the unit cost for electric power for the entire

year to the maximum electric power demand over a specified interval.

The interval might be several hours or longer during periods of heavy

usage. With this sort of clause, even a single day of excessive

electrical demand could result in significant increases in the total

annual cost of lighting energy. Higher peak cooling loads also require

larger equipment sizes to maintain comfort conditions, resulting in

higher first costs.

Thus, both energy efficiency and cost considerations stand as motivating

factors for the conduct of this research project. The objective of this

research is to determine energy transfer from lighting systems to

building spaces and thermal equipment as influenced by typical operating

conditions and equipment configurations, and to develop procedures to

promote the design of efficient lighting and HVAC systems, leading to

energy and cost savings. A combination of detailed full scale

measurements of lighting and HVAC performance and related computer

simulations forms the basis of the approach to these issues. This

interim report describes the results obtained to date on this ongoing

research effort. A final report will be forthcoming following

completion of the work.

1
2 . Background

The need to conserve finite energy resources has led to increased

efforts to maximize the efficiency of lighting, heating and cooling

systems in buildings. A major step in this regard has been the

integration of the light sources, or luminaires, with the air

distribution system[Be77, Bo77, Fi76, We71, No69] . Using supply or

return air to cool a fluorescent luminaire can raise lamp efficacy,

defined as the amount of visible light produced per unit of electric

energy consumed (and dissipated into the building space). In fact, both

the light output and power consumption, and thus efficacy, of a

fluorescent lamp vary with lamp wall temperature. Fluorescent

luminaires currently dominate the commercial building sector, and are

the focus of this study.

There are two major issues associated with the interaction of lighting

and HVAC systems in buildings. One issue is the energy performance of

the combined system. Energy for lighting will be minimized if the

luminaires are operated at their optimum temperature. The total heat

gain to the building space from the luminaire would also be minimumized

at this condition, meaning minimum cooling loads from lighting.

Luminaire temperatures are determined by the total room thermal

environment, including air and surface temperatures, and airflow rates.

The second issue relates to the peak cooling loads due to lighting.

When a lighting system is switched on, all of the electrical power input

does not show up as cooling load immediately, because of heat storage in

building components, plenum air (if any) and the luminaires themselves.

Of course, once the lights are turned off, the stored heat will be

dissipated to the room air and eventually appear as cooling load. The

result is the same total cooling load, but a lower peak load. Thus,

peak cooling loads due to lighting can be controlled by channeling some

of the energy dissipated by the luminaire into components such as floor

and ceiling slabs, walls and furnishings, thereby redistributing the

cooling load due to lighting over a longer period of time.

The capability of storing heat from the lighting system in the room

components is a function of luminaire type and location, airflow

configuration and thermal storage characteristics of available storage

elements. Massive, heavyweight building constructions can store more

energy, thereby moderating peak cooling loads to a greater extent.

The key to taking advantage of the thermal interactions between the

lighting system, the HVAC system, and the building to control energy

usage and peak cooling loads lies in understanding the energy

distribution from the lighting both under steady and fluctuating

operating conditions. The first step is to examine the temperature

dependence of the fluorescent lamps

.

2 . 1 The Temperature Dependence of Fluorescent Lamps

The popularity of the fluorescent luminaire can be traced to the

relatively high luminous efficacy (approximately 70 lumens/watt) of

fluorescent lamps, their large surface area, cool operation, and long

life[Lu38]. The fluorescent lamp is an electric discharge source

consisting of a glass tube containing mercury vapor at low pressure(~6-

2
10 pm of mercury) and an inert gas, such as argon at a higher pressure

(~2-3 mm) of argon [Mu85] . Light is produced when current flows between

electrodes at opposite ends of the tube creating an arc through the

mercury and generating ultraviolet radiation, which in turn is

transformed into visible radiation by fluorescent materials (phosphors)

on the inner tube walls. The primary wavelength of energy generated by

a mercury discharge is 253.7 nm. The phosphors are selected to respond

most efficiently at that wavelength. Fluorescent lamps must be operated

in series with a current limiting device called a ballast, which also

provides the required starting and operating lamp voltages.

The performance characteristics of fluorescent lamps are extremely

sensitive to temperature conditions, due to the effects on mercury vapor

pressure within the lamp. A fluorescent lamp contains more liquid

mercury than will become vaporized under typical operating conditions.

Any excess mercury tends to condense at the coldest spot on the lamp

wall. Thus the mercury vapor pressure varies with the minimum lamp wall

temperature. Changes in mercury vapor pressure affect both the light

output and power consumption of a fluorescent lamp

.

Maximum light output and power consumption occur when the lamp cold spot

is approximately 100°F, although they do not necessarily occur at

exactly the same temperature, as shown in figure 2.1. Maximum lamp

luminous efficacy also occurs near 100°F. The exact temperature

dependence of different lamps varies somewhat, but all display the same

general behavior as shown in figure 2.1[Ie72].^ As lamp wall

temperature decreases, light output and power consumption drop due to a

suppression of the mercury arc. The result is a one percent decrease in

efficacy for each 1 . 5°F decrease in lamp wall temperature. As lamp wall

temperature increases above 100°F, more mercury vaporizes, the mean free

path decreases and the proportion of elastic collisions increases.

There is also more extinction of radiation. The net effect is a one

percent decrease in efficacy for each 2.5°F increase in lamp wall

temperature [Mu8 5 ]

.

Other factors also influence fluorescent lamp performance including lamp

length and diameter, lamp loading and argon pressure. However, these

other factors are usually fixed for a particular lamp installation, and

not dependent on the thermal conditions. On the other hand, the thermal

conditions to which a fluorescent luminaire, that is the lamp and

fixture combination, is exposed can significantly alter lamp

temperature, and thus, lamp performance. These thermal conditions

include air temperatures, airflow conditions near the luminaire and

thermal radiation exchange with surrounding surfaces. It is important

to note that the thermal environment is somewhat under the control of

the building designer, allowing it to be tailored to promote efficient

luminaire operating conditions.

2.2 Energy Distribution From Fluorescent Luminaires

All of the power input to the lighting system eventually is dissipated

into the building space. About 15 percent of the power input to the

^-The figures for each chapter are located at the end of each

chapter

.

3
luminaire is consumed by the ballast before reaching the lamp. The

balance of input power goes to the lamps. Under typical steady state

conditions, about 21 percent of the input power to a fluorescent lamp is

radiated as visible light, while 37 percent is infrared radiation and 42

percent convective exchange with the air surrounding the luminaire

[Ie72] . The exact values of these energy distribution fractions vary

with the thermal conditions

.

When the lighting system is initially energized, some of the input

energy is stored in the luminaire. A fluorescent luminaire consists of

four main components, as shown in figure 2.2. In addition to the lamps

and ballast, these are the metal housing and the lens or diffuser. A

luminaire lens is usually fabricated from acrylic plastic with a

prismatic textured design to distribute the light from the lamps. In

contrast, a diffuser is an open framework of metal or plastic,

frequently with large parabolic cells to direct the light from the

lamps. A luminaire with an acrylic lens creates an enclosed lamp

compartment, while a parabolic diffuser allows air to circulate freely

between the lamp compartment and the surrounding environment.

As mentioned earlier, the light output from the lamps is a function of

minimum lamp wall temperature. Once steady luminaire temperatures are

achieved, which may take several hours, the remaining input power is

dissipated to the surroundings by infrared radiation exchange and

convective or conductive heat transfer. The convective fraction is

dependent on airflow conditions at the luminaire. In some cases, the

ventilation system is utilized to circulate air through the lamp

compartment, or through slots at the perimeter of the luminaire, to help

cool the lamps by increasing the convective heat transfer [Be77, Bo77,

Mu62] . Thus the airflow rate and configuration affect the radiant and

convective heat dissipation fractions.

Frequently, lighting systems in commercial buildings are in a recessed

mounting above a suspended ceiling, in a plenum space. In that

configuration, energy transfer can occur to both the room below the

suspended ceiling, and the plenum above. The plenum may also be used to

channel return air from the room to the circulation fan duct inlet.

Table 2.1 summarizes the various energy transfer fractions from a

lighting system to the surrounding environment, including the room,

ceiling and plenum.

Table 2.1 Energy Distribution Fractions from Lighting System

1. Visible light to room

2. Visible light to plenum

3. Convection to room air

4. Convection to plenum air

5. Convection to return air

6 . IR radiation to room

7. IR radiation to plenum

8. Conduction to suspended ceiling

9. Stored in luminaire

Figure 2.3 displays these energy distribution fractions schematically.

The values of each of the fractions influences the cooling load profiles

due to transient or cyclic operation of the lighting system. Knowledge

4
of these fractions is necessary to accurately determine cooling loads

for equipment design purposes. As will be described later in this

chapter, building energy analysis computer programs require the values

of these fractions as input parameters

.

2.3 Cooling Load Due to Lighting

There are several factors which influence the cooling load due to

lighting. The luminous efficacy of the lighting system determines the

amount of heat dissipated into the building space to provide the

required light levels. Less efficient lighting systems use more energy

to provide the same, light levels; consequently, more heat must be

removed from the building space to maintain comfort conditions. Cooling

load profiles due to cyclic operation of the lights are also dependent

on the thermal storage characteristics of the room and plenum.

It has been observed that the cooling load due to a step change in

lighting power can be approximated by an exponential relation[Ki68 ] . If

the lighting system is suddenly energized in a room at thermal

equilibrium, the cooling load Q is related to lighting power W according

= i - Ae'

Bt

(2.1)

w

where: t = time

A,B = curve fitting constants

Figure 2.4 displays this relation graphically. The values of A and B

can be determined from measurements, or detailed calculations. Equation

2.1 approximates the cooling load when the lights are initially turned

on by assuming an instantaneous cooling load equal to 1-A. This

represents heat dissipation from the lighting system directly into the

return air. In actuality, the initial rise in cooling load is steep but

requires a finite time interval to reach the value of 1-A.

If the lighting system is operated on a cyclic (i.e. daily) schedule and

the room has any appreciable thermal mass, the cooling load due to

lighting will not yet equal the total lighting power when the lights are

turned off. All or part of the heat from the lights which had been

stored in the room components will be discharged to the room air, thus

contributing to cooling load, as long as the circulation fan and cooling

system are in operation. This general effect is shown in figure 2.5.

For an interior office space in which no heat is lost through the

exterior envelope, the total daily cooling load is the same whether or

not any heat storage occurs. Peak cooling loads, however, can vary

significantly depending on the heat storage characteristics of the room

and plenum. In most buildings, all of the heat stored in the thermal

mass of the building space will not be reclaimed during the period in

which the lights are turned off. Thus when the lights are turned back

on the next day, less storage capacity is available, and cooling loads

are higher. Eventually, after a number of cycles, the cooling load

profile reaches a steady periodic condition, and the actual peak cooling

load is determined.

5
2.4 Weighting Factors and Design Procedures

Building designers need to be able to predict cooling loads in order to

determine equipment sizes, both for fans and cooling equipment. There

are many heat sources which contribute to cooling loads in buildings,

including lighting, equipment, solar gains and envelope heat gain. The

usual approach to determining total peak cooling loads has been to

determine the cooling load profile due to each source separately, and

then to combine them to determine the peak concurrent cooling load. One

of the most widely used methods of achieving this involves weighting

factors

.

Weighting factor techniques result from the application of z- transform

procedures to transient heat transfer phenomena. The z- transform is, of

course, useful for describing processes which are evaluated at discrete

time intervals, such as sampled-data systems. In this application,

cooling load is usually calculated at intervals of one hour. In this

manner, a continuous cooling load profile, such as that given in

equation (2.1) is transformed into a sequence of hourly values. The

cooling load response due to a series of lighting step changes, either

off or on, can be calculated once the weighting factors are known.

An exponential relation of the type given in equation (2.1) (i.e. 1 - A

exp (- Bt) leads to a discrete time series relating cooling load to the

time history of cooling load and lighting power, according to the

following derivation.

Equation 2.1 represents the output (cooling load) response to a step

change in input lighting power. Thus, in z space

G(z) is simply the function which relates input to output. A unit step

change in lighting has a z- transform of

Q(z) = G(z) W(z) ( 2 . 2 )

where: G(z) = z- transfer function

z = e Ts

T = time increment

s = Laplace variable

(2.3)

while the z- transform of the output is

Az

(2.4) -BT

z - e

Solving for G(z) we obtain:

2

Az

-BT

(2.5)

- ze

6
Combining terms , equation 2 . 5 leads to

G(z)

(1 - A)z° + (A - e~

BT

)z~

1

1 - (e ) z

( 2 . 6 )

The weighting factors are defined by the coefficient of the z terms in

equation 2.6, as follows:

aQ = 1 - A

ai = A - e b} = e-BT

-BT (2.7)

Thus equation 2 . 6 can be written

o -1 ... a z + a.z

Q(z) _ o 1

><> i

.

( 2 . 8 )

Finally, transforming equation 2.8 back into the time domain produces

Q(t) = aQW(t) + a X W(t-T) + bxQ(t-T) (2.9)

Equation 2.9 allows the calculation of cooling load at any time step

based on the current lighting power and the lighting power and cooling

load from the previous time step.

While equation 2.9 corresponds exactly to equation 2.1, in reality

equation 2.1 does not accurately reflect cooling load during the first

two or three hours of a step response test[Mi74] . This is due to the

presence of other short term transients. According to equation 2.9, an

instantaneous cooling load occurs, equal to 1-A, as soon as the lights

are energized. This overestimates cooling load during the initial time

steps. One technique to improve the accuracy of the cooling load

prediction is to drop the a0 term in equation 2.9 and replace it with an

a2 term [Mi74] . While the resulting relation no longer corresponds

exactly to equation 2.1, it was judged to be more representative of the

actual cooling load profile. The adjusted relation is:

Q(t) = a x W(t-T) + a2W( t-2T) + b^t-T) (2.10)

The values of the weighting factors for equation 2.10 are as follows:

a 1 = Q(l)/W(l) = 1 - Ae‘ B ( 1 )

7

( 2 . 11 )
where: the substitution t = 1 has been made

The value of a£ results from the fact that at t = «, cooling load equals

lighting power

Equations 2.10 and 2.11 are the basis for the cooling load due to

lighting calculation procedures found in the ASHRAE Handbook of

Fundamentals [As77] . Values for the weighting factors are listed for

various room constructions, lighting system types and airflow rates.

Also, cooling load factors (CLF's) are listed in tabular form for

various lighting daily cycles (i.e. 8, 10, 12 hours on). These were

determined directly from the weighting factors.

The weighting factors themselves were determined from measurements and

analytical procedures . A series of measurements were carried out at the

National Research Council Canada [Mi71, Mi73, Mi74] , along with some

transient thermal modeling [Mi78, Mi83]. The basic measurements

consisted of monitoring the cooling load due to a step change in

lighting in a room size calorimeter. The values of the weighting

factors were determined from the values of A and B in equation 2.1. The

A and B values can be determined as follows. For a lights - on test:

sSLO 1! -Bt

- Ae (2 . 1) (2 . 12)

ln [l - ^1 = In A - Bt (2.13)

For a lights - off test:

9 = Ae"

Bt

W

A

In m- In A - Bt

(2.14)

(2.15)

Equations 2.13 and 2.15 enable the determination of A and B using least

squares linear regression, where:

Y = a + bt (2.16)

and

Y = In

(

1 - I )or ln

( I )

a = In A, A = e a

b = -B, B = -b

2.5 Building Energy Analysis Computer Programs

With the advent of modern digital computers, building designers have

added new tools to their arsenal of techniques for evaluating building

performance. In particular, computer programs have been developed to

calculate building energy usage for hourly or other periodic time

8
increments. These programs calculate cooling, heating and lighting

loads for transient heat storage, and heat gains due to building

envelope heat transfer, internal equipment and occupancy factors. Two

of the more widely utilized programs are BLAST (Building Loads Analysis

and System Thermodynamics) [Hi79] and DoE-2 [Do79]

.

These two programs fall into two general categories according to their

basic solution procedure. BLAST, and other similar programs, are based

on heat balance calculations. Following this technique, the building

space is divided into elements which can communicate by conduction,,

convection, or radiation. The sum of all heat transfer to each element

is computed for each time step. Any net heat gain results in an

increase in temperature of the element, or a cooling load if the nodal

temperature is held constant. A net heat loss causes a heating load.

Computer programs like DoE-2 use weighting factors to determine building

energy usage. Separate weighting factors are used for each heat source,

such as lighting, electric equipment, solar heat gain and exterior

thermal heat transfer. Tables of weighting factors are stored for

various constructions or custom weighting factors can be computed when

needed.

For either type of computer program, the simulation of lighting system

performance and related heat gains is accomplished by using user input

values. Constant values are input by the program user regarding maximum

lighting power and the lighting energy distribution fractions, such as

percentages of lighting input power which reach the room as visible

radiation, infrared radiation or convection, or enter the return air

directly. These values are treated as constants for the duration of a

simulation. A lighting schedule is also specified which determines the

fraction of the lighting power which is energized for each hour. This

accounts for switching of the lights due to occupancy patterns.

The use of constant values for the lighting energy distribution

fractions may be adequate for many applications of BLAST and DoE-2.

However, there are some drawbacks to this approach which may prove to be

substantial in many instances. First, the selection of the appropriate

values for the fractions is difficult since the fractions are dependent

on the thermal conditions and room configuration. Thus, they are

dependent variables. Limited data are available to assist the selection

of the fraction values. Second, the lighting energy fractions are not

constants, but vary with the thermal conditions, especially airflow

rates and temperature. The use of constant fractions means that the

distribution of lighting energy is assumed to be fixed regardless of

operation of the HVAC system. Third, maximum lighting power is assumed

to be constant, disregarding the lamp temperature dependence. As a

result, the calculation of cooling load due to lighting is subject to a

uncertainty of five to fifteen percent at best, and even greater

uncertainty if inappropriate lighting energy fraction values are used as

input parameters.

The lack of sensitivity of BLAST and DoE-2 to lighting and HVAC

interaction effects minimizes the usefulness of these programs for

evaluating and comparing different lighting and HVAC system design

alternatives. One approach to alleviating this deficiency would be to

change the lighting energy fractions from constant values to variables

9
which are functions of other simulation variables. The most logical

variables would be room air temperature, mean room surface temperature

and airflow rate.

2 . 6 Summary of Previous Work

Lighting system performance and its relation to room thermal conditions

has been a subject which has attracted attention since the development

of fluorescent lamps. The initial emphasis was on ensuring adequate

light levels at the expected operating conditions . With the advent of

mechanical cooling, the determination of cooling loads due to lighting

became a subject of interest. More recently, increasing energy costs

and limited electrical power generating capacity have led to increased

interest in the development of efficient lighting systems and the

control of peak cooling loads due to lighting.

One of the earliest references to removing the heat from lighting [Ei33]

dealt with combination incandescent and mercury lamps, which dissipated

more than 15w/ft^ of floor area. The lamps were enclosed in a heatabsorbing glass cylinder and mechanically ventilated to the exterior at

an air flow rate of 600 ft^/min in warm weather. The exhaust air was

recirculated to the building during the heating season. The lighting

power density of 15 w/ft^ was very high by modern standards.

With the introduction of the fluorescent lamp in the late 1930' s and

early 1940' s, much lower lighting power densities were capable of

providing the same light levels [Lu38] . The effects of fluorescent

lamps on cooling and comfort were addressed in [Da40] and design factors

for lighting and air conditioning were published in [He41]. Lighting

was identified to account for an average of 18 percent of total cooling

load, and to range from 2 to 47 percent for different installations.

Recommendations for fluorescent lighting practice were published in

[Ed39]„ These papers and reports were based on limited measurement data

and basic theoretical calculations, and were not validated for a wide

range of conditions. However, they did indicate an appreciation for the

thermal effects of lighting even at the early stage of fluorescent

lighting system development.

The recommended procedures for designing fluorescent lighting systems

and estimating cooling loads due to lighting changed very little through

the 1950' s. The Illuminating Engineering Society published a guide to

lighting and air conditioning [Ie56] which recommended assuming 100

percent of lighting power becomes an instantaneous cooling load. This

recommendation was justified by lack of data on heat storage effects,

and represents a conservative cooling load estimate leading to oversizing of the cooling equipment.

The lack of data on the thermal performance of fluorescent lighting

systems led to several research efforts in the early 1960's. The use of

small calorimeters to measure the heat pickups by air circulated through

a luminaire became popular. Several researchers used this approach to

estimate the energy distribution characteristics of various types of

luminaires [Mu62, Ba63, Ba64, Ba65, Sy64, Fi75, Ne71]. Typically, the

calorimeters consisted of large insulated boxes, with the luminaire

mounted on the bottom face. An external fan attached to a duct allowed

air to be drawn through the luminaire into the box and then exhausted

10
from the box, which served as a plenum. The boxes were usually

instrumented with temperature sensors, and luminaire light output and

power consumption were measured, along with ventilation air flow rate.

These luminaire calorimeters provided much needed information regarding

the thermal performance of lighting systems. The temperature dependence

of lamp light output and power consumption was demonstrated.

In spite of the usefulness of luminaire calorimeters, they had several

weaknesses which did not allow them to address all of the lighting

thermal performance issues. The primary element lacking from the

luminaire calorimeter was the room in which the luminaire would be

installed. Because the luminaire calorimeter did not emulate a room

with a plenum, the thermal conditions surrounding the luminaire were not

representative of a typical room environment. Heat transfer by

radiation, and heat storage in the room and plenum could not be

evaluated. Also, convective heat exchange would likely be different in

an actual installation than in a small calorimeter. However, design and

analysis techniques were published based on calorimeter data [Me63,

Me64, Qu62, He66] . Also, field studies were undertaken to demonstrate

the application of design procedures [Du64a, Du64b , Fi66].

The Illuminating Engineering Society developed an approved method for

the photometric and thermal testing of fluorescent luminaires [Ie68].

This described a calibrated calorimeter and specified a testing and data

analysis procedure. This method has been widely utilized.

Progress in the characterization of the thermal performance of lighting

systems continued with full scale room tests of fluorescent luminaires

[Bo66a, Bo67, Bo68a, Bo68b, No69]. The first four of these references

were primarily concerned with comparing steady state temperature and

airflow conditions with different air handling systems. The latter

reference was concerned with heat transfer from the luminaire to the

room and plenum under steady state conditions. None of the studies

addressed the transient lighting thermal performance or heat storage

issues

.

The first transient full scale room lighting thermal performance study

was conducted at the National Research Council Canada. Following an

initial analytical study [Ki68], a full size room calorimeter was

constructed and instrumented [Mi71, Mi74] . The test facility consisted

of a test room with floor and ceiling slabs, with a plenum and lighting

system in the room and below the floor. The test room walls were foilcovered foam insulation, and the entire structure was enclosed by an

insulated shell. Room temperature air was circulated between the room

walls and the exterior shell. The temperature of the top of the ceiling

slab was controlled by a pond of water at the same temperature as the

floor surface. All of the fans and heating and cooling equipment were

located in the test room. Cooling load due to the lighting system was

determined through an energy balance relation using the temperature rise

of the cooling coil water and the measured heat addition and fan power.

The return air temperature was held constant by varying the amount of

heat addition from an electric heater.

The NRC test facility was used to measure cooling load profiles for a

variety of configurations, including various supply and return air

locations, with and without furnishings and carpet. Weighting factors

11
were calculated as described earlier. The availability of these data

represented a great stride forward for the calculation of cooling loads

due to lighting, and the weighting factors were incorporated into the

ASHRAE handbook of fundamentals [As77] . Additional weighting factors

were computed analytically [Mi73] . Methods for utilizing the weighting

factors are described in [Mi72, Mi78, Mi83]

.

The other aspect of previous lighting thermal performance research

relates to computer simulation of lighting systems in building spaces.

Due to . the limitations in the number of configurations which can be

measured, computer models are useful for extending the generality of the

results. Two major lighting system modeling efforts have been

published. The first computer model focussed on steady state

performance using network analysis [So72, So73, So74] , drawing on the

measurement results from [No69] . The main emphasis was on temperature

distributions, comfort conditions and the fractions of lighting energy

transferred to the room and plenum. The effect of different return

airflow rates and luminaire conditions was also simulated. No transient

simulations were performed.

More recently, a detailed transient lighting thermal performance model

was been developed [Ba83a] . This model was based on heat balance

calculation procedures in an explicit transient finite difference

formulation. Cooling load profiles were calculated and compared to some

of the results from [Mi73]. The model used empirical correlations for

determining convection coefficients, and accounted for the temperature

dependence of light output and power consumption. Based on a series of

simulations, weighting factors were derived and published [Ba83b]

.

To summarize the current state of the art, cooling load profiles are

computed using ASHRAE- type weighting factors based on a limited set of

measurements and simulations. Very little information is available

regarding the lighting energy distribution fractions which are needed as

input parameters for building energy analysis computer simulations. No

data are available regarding heat storage in room components due to

cyclic operation of the lighting system.

2.7 Objective of the Present Research

The objective of the research reported here is to increase the

understanding of the interactions between lighting and HVAC systems in

commercial - type buildings. Specifically, the thermal performance of

fluorescent luminaires under typical office conditions will be evaluated

based on detailed full - scale measurements and related computer

simulations. The primary elements of the research are as follows:

o Measure cooling load profiles due to lighting for various room

temperatures, airflow rates and airflow configurations.

o Determine ASHRAE - type weighting factors from measured cooling load

profiles

.

o Investigate the possibility of using another time function to fit

calculated cooling loads to measured loads more exactly than the

single exponential relation which forms the basis of the ASHRAE

weighting factors.

12
o Determine extended weighting factors to correspond to improved

cooling load profile function.

o Examine heat transfer processes from lighting and determine lighting

energy distribution fractions.

o Examine heat storage in room and plenum components.

• o Examine air and surface temperature conditions

.

o Evaluate effect of room air temperature, airflow rate and airflow

configuration on lighting and HVAC system performance.

o Develop, calibrate and validate a detailed computer model of the

lighting and HVAC interaction test facility.

o Suggest design and operation strategies to control peak cooling

loads due to lighting and total energy for lighting and cooling.

13
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Ballast

Figure 2.2 Schematic Diagram of Fluorescent Lamp Components

15
Ceiling Slab

Figure 2.3 Energy Distribution Fractions from Lighting Systems

16
COOLING

L

CUD/

LIGHTING

POWER

Figure 2.4 Cooling Load due to Step Change in Lighting

COOLING LOAD STEP RESPONSE

7V-t.H0URS COOLING + LIGHTING

COOLING LOAD WITH CYCLIC LIGHTING

Figure 2.5 Cooling Load with Cyclic Operation of Lighting System

17
3. Experimental Facility and Plan

The lighting and HVAC interaction test facility was designed and

constructed to emulate an office space using standard materials and

construction techniques. The test facility is extensively instrumented

to allow the measurement of cooling load, lighting power, airflow,

surface and air temperatures, heat flows and light levels. Three air

handling units service seven individually controlled zones, including

the test room and surrounding spaces. Data collection and control are

accomplished using a computer-controlled system.

The test facility is designed to allow testing configurations to be

changed easily. Airflow rate, air return path and room temperature are

three of the primary testing parameters. The unique features of the

test facility are the guard air spaces, which simulate adjacent building

spaces, and the lower plenum with duplicate lighting and HVAC system

beneath the test room floor. The air circulation system for the test

room emulates a typical building design.

Additional measurement parameters to be evaluated include various

luminaire types, number of lamps, the effect of carpet, different wall

types, the effect of variable supply air volume and the impact of an

exterior condition adjacent to a perimeter wall.

3.1 Test Facility

The test facility is constructed on a slab 30 ft. 6 in. by 21 ft. 4 in.

within the large NBS environmental chamber. The facility is divided

into two sections, a large insulated shell enclosing the test room area,

and a smaller attached control room for housing instrumentation. The

overall height is 20 ft. 10 l/2in. , while the control room ceiling

height is 13 ft. 2 1/2 in. The test room floor slab is elevated to

accommodate a lower plenum beneath the floor, and all other room

surfaces are adjacent to temperature - controlled guard air spaces.

Duplicate lighting and HVAC systems are installed in both the test room

plenum and the lower plenum. Figure 3.1 shows a cut - away schematic

view of the test facility. The majority of the ductwork is not shown in

this figure

.

The test room floor and ceiling slabs are 2 1/2 inch thick concrete

built on steel decks supported by a structural steel framework, as shown

in figure 3.2. The test room walls are constructed of gypsum drywall on

steel studs. The initial test configuration is four interior walls.

The guard air spaces are shown in figure 3.3.

The lighting system locations in the test room and lower plenum are

shown in figure 3.4. The edges of the floor slabs extend slightly

beyond the walls, while the ceiling slab separates the side guard air

spaces from the upper guard air space. Small access doors allow entry

into the guard air spaces , as shown in figure 3.5. Also shown is a

large double door which opens from the north guard air space. Another

door connects the test room/guard air portion of the test facility with

steps leading to the floor of the control room.

18
Chilled water for cooling, and electrical power are drawn from

environmental chamber supplies. The temperature of the chamber is

maintained at approximately 75°F.

During and following construction, the test room, guard air spaces and

ducts were sealed and insulated where appropriate to minimize unwanted

heat and air flows. All doors were gasketed and joints sealed. All

conduits through the test room walls are made of plastic to reduce heat

flow.

Each air handling system consists of a fan and chilled water coil, with

separate electric duct heaters for each supply duct. The output of each

of the electric duct heaters is controlled by a separate 4 to 20 mA

control signal, allowing independent temperature control for each of the

seven supply air systems. The test room air system is dedicated to the

test room only. Return air is taken from the plenum near one corner of

the room, through a variable speed fan, over a cooling coil, past an

electric duct heater and supplied through diffusers at suspended ceiling

height. All ductwork is insulated. The air temperature leaving the

cooling coil is controlled to set point by an automatic valve which

varies chilled water flow. Room air temperature is controlled by

varying the electric power to the duct heater. Figure 3.6 shows the

test room ducts.

The second air handling system services the east, south, west and upper

guard air spaces. Chilled water flow is adjusted manually to keep air

temperature after the cooling coil below the minimum required for

testing, so that control is maintained by modulating the control of the

heater. Figure 3.6 shows the east, south and north guard air duct

systems. Supply air is introduced at the top of the vertical guard air

spaces, and extracted at the bottom. In the upper guard air space, air

is supplied at the center and extracted around the edges.

The third air handling system services the lower plenum and north guard

air space. The north guard air space is designed to function as either

an interior or exterior condition, although was only used as a interior

space during these tests. The lower plenum supply duct is identical to

the test room, while the return air inlet is in the opposite corner.

Figure 3.7 shows the duct layout in the lower plenum, and air handling

unit number three.

The electrical system design is standard, except for sub-metering of the

lighting, auxiliary electric power, fan power and test room heater

power. The seven electric duct heaters and variable speed fan drive are

individually controlled by the computer system, as will be described

later. Identical lighting systems are installed in the test room and

lower plenum.

3.2 Instrumentation

The operation of the test facility, data collection and control of

measurements is by a personal computer based data acquisition and

control system. The system consists of an IBM AT and three Keithley

Series 500 units. The computer collects data and controls the heaters

and fan and other measurement parameters under the control of a

19
specialized computer program written in SOFT500, an extended version of

BASIC. A total of 398 parameters are recorded at two minute intervals.

All channels except the surface temperatures are sampled every 12

seconds, and then values averaged over two minutes. The 12 second

readings for the photometers are an average of 100 samples to negate the

effects of flicker from the fluorescent lamps.

One day's worth of data takes 2.2 mbyte of storage. Following a test,

which may last several days,, data is stored on a cartridge storage

device and transferred to another IBM AT computer for analysis. The raw

data is first loaded into a database system called RBASE and then

subsets of the data are exported to LOTUS for processing, plotting and

analysis

.

The raw data is also analyzed by a program which computes average values

for 30 minute intervals. Computed averages include floor and ceiling

temperatures, wall temperatures, air temperatures, light levels and heat

flows

.

The primary measurement parameters are lighting power, cooling load,

return airflow rate and room temperature. Additional measurements of

air and surface temperatures, heat flows and light levels are intended

to supply supporting information. Some of the measurements also serve

as feedback signals for the control loops, such as the floor and wall

thermopiles

.

t Lighting power is measured using a solid state transducer with a current

output proportional to power consumption. This sensor was calibrated at

NBS and found to have an uncertainty of less than 0.1 percent.

The sensible cooling load is derived from the temperature difference of

the air from the room supply to return, and the airflow rate. The

return airflow rate is measured using an array of four hot wire

anemometers in the return air duct. These sensors measure mass velocity

and do not need to be corrected for density. The airflow rate in cfm

(AF) is given by:

ft 2

AF = (Vel) — (duct area) ft (conversion factor) (3.1) mm

The conversion factor is necessary to account for the true volume flow

corresponding to the measured velocities. Since airflow velocity varies

through the duct cross - section, the measured velocity may differ from

the average. The conversion factor was determined by comparison to the

airflow measured with a detailed traversing of the return air duct. The

duct area is 0.92 ft^. The airflow velocity uncertainty is 3 percent.

20
The cooling load for the test room is given by:

Q = mCpAT = (AF) (p) (Cp) (Tret ' TSUP) (3.2)

where: p — density of standard air

Cp = specific heat of air

Tret = return air temperature

Tsup = supply air temperature

The cooling load in watts is:

Q (watts)- (AF) ff. (,) ^3 (Cp) (4T) F

" 1 w

3.413 Btu

hr = (AF) (0.08056) ——r (0.240) (4T) F (17.580) min - lD*r

ft = (0.33990) (AF) (AT)

(3.3)

The supply to return air temperature difference is measured using a type

T thermopile with 42 pairs of junctions. At 23 /xv per °F per pair of

junctions, this thermopile generates 1 mv per °F temperature difference.

Although slightly non - linear in response, throughout the range of

testing the non - linearity is less than a one half percent effect, and

uncertainty is approximately 0.1°F.

The test room air temperature is measured with an array of 64

thermocouples, in a four by four by four grid, as shown in figure 3.8.

The upper 16 thermocouples are in the plenum, while the average of the

48 thermocouples below the suspended ceiling is used as the room

temperature control point. Temperature measurement uncertainty is

0 . 75°F.

Surface temperatures are measured throughout the test room, primarily on

the north, east and south walls, floor and ceiling, as shown in figure

3.9. Two of the luminaires have 16 thermocouples each, while the other

two luminaires have 8 thermocouples each.

Heat flux transducers are mounted on each room surface, at the center

and around the perimeter, as shown in figure 3.10. The center

transducers are one square foot in area, producing a miiivolt signal

proportional to heat flow through the surface. The other heat flow

sensors are smaller disks or squares. The heat flow meters were

calibrated using the NBS Guarded Hot Plate. Uncertainty is

approximately 5 percent

Thermopiles are used extensively to control the test room boundary

conditions. The most typical boundary condition is to simulate a test

room surrounded by similar spaces. Thus, wall temperature conditions

would be symmetric about the center plane of each test room wall, and

the surface temperature would remain equal on both sides of the wall.

The average temperature difference across each wall is measured using

Type T thermopiles with 30 pairs of junctions. The control system

21
attempts to keep the thermopile readings equal to zero by varying the

power supplied to each electric duct heater. When the lights are

switched on, the interior of the test room walls begin to heat up,

causing an imbalance in surface temperatures across the walls which is

sensed by the thermopile which, in turn, causes the control signal to

the appropriate heater to increase. In this manner, heat can be stored

in the gypsum board walls, but no net heat flow through the wall will

occur. Figure 3.11 demonstrates the wall thermopile system

schematically

.

The control of the lower plenum and upper guard air space is also

accomplished using thermopiles. The intent is to maintain the

temperature at the top of the ceiling equal to the top of the floor, and

the bottom of the floor equal to the bottom of the ceiling. Net heat

flow will occur through the floor and ceiling, but should be equal in

magnitude and direction. The temperature distribution will vary across

the floor and ceiling slabs due to the presence of the lighting system.

Duplicate lighting systems below the floor and ceiling provide

equivalent temperature distributions, while the thermopiles control

average temperatures to equivalent levels. One 36 pair junction

thermopile is attached to the top of the ceiling and floor slabs and

another is attached to the bottom of the two slabs. Power to the

electric duct heaters is varied to maintain equal surface temperatures.

The control loops operate on a 12 second sampling interval. A digital

proportional and integral control algorithm is implemented to vary

electric heater power as required in response to the thermopile

readings. The change in controller output, m, is related to the error

signal, e, according to [Ra78]:

Am = m(K) - m(K-l) = K

p

[e(K) - e(K-l)] + KjTe(K) (3.4)

where: K = sample

T = sampling interval

Kp= proportional gain constant

Kj= integral gain constant

At the start of a test, the heaters are all off and are gradually

increased automatically to achieve and maintain the desired condition.

Values for K

p

and Kj were determined by trial and error.

Light levels in the test room are monitored continuously at the center

of each surface using silicon photovoltaic cells corrected for eye

response. Also, periodically detailed light distribution measurements

are made using an array of photocells. The photocells were calibrated

using an NBS reference source, and have an uncertainty of less than one

percent.

In addition to the sensors described above, numerous other air and

surface temperatures are measured in the guard air spaces and air

handling systems to help monitor test facility performance.

3 . 3 Measurements and Calibration

Both steady state and transient tests can be run in the test facility.

Steady state tests involve establishing a test configuration and

22
allowing conditions to stabilize. The transient tests involve the

response of cooling load due to sudden switching of the lighting system.

Thus, the test facility goes from one steady state condition, either

lights on or off, to the other steady state condition.

The initial tests were used to calibrate the test facility, tune the

digital control loops and verify the proper operation of the equipment

and sensors. The test facility has two natural calibration conditions,

that with the lights off (zero cooling load) , and with the lights on

(cooling load equal to lighting power) . Initial steady state

measurements indicated excessive heat gains to the test room with no

lights on. Additional insulating and sealing of the test room envelope

alleviated this problem.

The air supply rates for the guard air spaces, and chilled water flows,

were adjusted to enable their proper temperature control. Air must be

supplied to the guard air spaces at the range of temperatures required

to match surface temperatures with the test room. Thus the air

temperatures after the cooling coils must be below that needed for the

guard air spaces, so that the guard air supply temperature can be

controlled through the addition of heat by the electric duct heater.

The airflow rate to each of the guard air spaces must be great enough to

control the surface temperature, but low enough to match the room

surface temperatures when they are heated by the lighting system. These

rates were adjusted by trial and error.

Test room air leakage was measured using sulfur - hexafluoride tracer

gas decay. Natural infiltration was found to be 0.017 air changes per

hour. Sulfur hexafluoride was injected into the test room with fans off

and ducts sealed. The decay in SFg concentration was monitored over

time and the air infiltration (I) calculated from:

where : t = time interval

CQ= initial SF5 concentration

C = final SFg concentration

Sulfur hexafluoride concentration was determined using an electron

capture detector.

The values for the gain constants for the digital control systems were

determined by monitoring heater operation and cooling load in response

to a step change in lighting. Initially, Kp and Kj were set to unity.

Excessive oscillations in cooling load and heater operation were

observed, as shown in figure 3.12 and 3.13. The gains were successively

reduced until oscillations were minimized, while sufficiently fast

response was maintained. Gain values of 0.1 were found to be most

effective

.

One of the key indicators of the performance of the control systems is

the thermopile output values during a transient test. The thermopiles

should read zero if the control systems are operating properly. Figure

3.14 shows the outputs from the supply return air thermopile and one

wall thermopile. The lights were switched on at hour seven, causing the

supply to return air temperature difference to increase as cooling load

23
picks up. The wall thermopile, however, maintains an average zero

reading, with some small, rapid fluctuations due to air turbulence and

controller operation. The other thermopile outputs are shown in figure

3.15 for the same test, on an expanded scale. The average temperature

difference as measured by the thermopiles is within 0.03°F of zero for

each sensor. The rapid temperature difference fluctuation are generally

less than 0.2°F. This performance was also observed for other tests.

3.4 Test Parameters

The primary testing parameters evaluated in this report are airflow

configuration, airflow rate and room temperature, and the testing is

aimed towards evaluating the effects of these parameters on the

transient and steady state performance of the lighting and HVAC system.

A series of tests was executed for 27 basic configurations, for three

variations of each of the primary parameters. The testing parameter

variations are summarized in table 3.1. This table also lists the code

for each parameter option. These codes appear throughout the figures

for identification purposes . Additional test parameters and room

configurations are scheduled for future testing. These are also listed

in table 3.1 and labeled with *.

Table 3.1 Test Parameters

Airflow return - ceiling grill (code CG)

luminaire side slots (code SS)

lamp compartment (code LC)

Airflow rate - 30 cfm per luminaire, 120 cfm total (code 120)

40 cfm per luminaire, 160 cfm total (code 160)

50 cfm per luminaire, 200 cfm total (code 200)

Room air temperature - 70°F, 75°F, 80°F

Luminaire type - parabolic diffuser, prismatic lens

Number of lamps per luminaire - 2*, 4

With/without carpet*

Wall type - 4 interior walls

3 interior walls, 1 exterior curtain wall*

3 interior walls, 1 exterior block wall*

Constant or Variable* supply air volume

Night - setback*

The luminaires tested were four, four- tube units with parabolic

diffusers. Each luminaire had two ballasts and four standard 40 watt

lamps. Return air could be directed either through the lamp

compartment, through slots on the sides of the luminaire, or through the

ceiling grill return. Additional tests will be performed using fourtube luminaires with prismatic lenses, and two-tube luminaires with

parabolic and prismatic diffusers.

24
Lighting/HVAC

Test

Facility

Figure 3.1 Cut-away Schematic View of Test Facility

25

Floorslab
Figure 3.2 Floor and Ceiling Slabs are Supported by Steel Frame

26
Figure 3.3 Guard Air Spaces Surrounding Test Room

27

PLAN

VIEW

OF

LIGHTING

TEST

FACILITY
Figure 3.4 Lighting Systems

28

LIGHTING

SYSTEM

PLAN
Figure 3.5 Access Doors Allow Entry to Guard Air Spaces

29
Figure 3.6

f? 8

^5 Z3 ± q. 2(3<w

Test Room Air Supply and Return Ducts, and Guard Air Ducts

30
Figure 3 . 7 Duct Layout in Lower Plenum

31

West

Guard

Air

Space
Figure 3.8 Test Room Air Temperature Thermocouple Grid

32
* • • • •

• • • •

North Wail • • • •

• • • •

1 •

1 • • • • •

”1 • • | •

1 1

1 • • • • • • • 1 •

| West Wail Floor East Wall • • • • • • • • ! •

1

1

1 • • • • •

1 • • | •

i

1

1 •

l • • • • •

1 • • 1 •

l • • •

• • • • Measurement Sta

South Wal

• • •

• • •

Figure 3.9 Surface Temperature Measurement Grid

33
South Wall

Floor

| East Wail #

i

West Wall j

North W*

• Heat Flux Transducer

- H I— ift

Figure 3.10 Heat Flux Transducer Locations

34
CONTROL

OF

GUARD

AIR

TEMPERATURE

3 O

H O O

<3 " W

/ ® «- ill

•i i £

1

• •

• •

• •

• •

• •

«

c

O

0

c

3

—3

©

1

i

©

(5

Figure 3.11 Wall Guard Air Thermopile Control System Schematic

35

Thermopile

output

Is

dc

voltage

(mv)

proportional

to

average

temperature

difference

across

wall
•otl*

or

c

fm

(Thouaonda)

Figure 3.12 Cooling Load Oscillations Due to Excessively High Control

Loop Gain Constants

Figure

cf231

8

bma.rrun O coo bgWinq © CMrflov

t 2B © 2C 4 2A X SB © .SA

.13 Heater Control Signal Oscillations Due to Excessively High

Control Loop Gain Constants

36
Figure 3.14 Outputs From Supply - Return Air Thermopile and Wall

Thermopile During Transient Test.

C

Q

Figure 3.15 Wall Thermopile Outputs on Expanded Scale

37
4. Computer Model of Lighting and HVAC Interaction

One of the objectives of this research is to determine the energy

distribution fractions from fluorescent luminaires under typical

operating conditions. Most of these fractions cannot be measured

directly. For example, while the distribution and amount of visible

light from the luminaires can be measured with photometers, heat

transfer from the luminaires by convection and infrared radiation

cannot. As a result, another method for determining the distribution

fractions is needed, which can be validated by the measurements.

The method chosen to determine the lighting energy distribution

fractions is a computer implementation of an analytical model. By

modeling the test facility, including the lighting system and room

components,, the lighting energy fractions can be computed. Calibration

of the computer model with the measured data verifies the accuracy of

the computed fractions.

One additional benefit of the computer model is that simulations can be

executed for more lighting and room configurations than can be tested,

due to limitations in time and physical constraints. Thus, the computer

model is useful for extending the measurement results to a larger

collection of configurations.

4.1 Basis of Computer Model

The computer model is a transient finite-difference formulation using an

explicit solution procedure. Heat balance equations are solved for each

node, with each node representing a major component of the room or

lighting system. The net heat transfer by conduction, convection,

infrared radiation and visible radiation is determined for each node.

Any net heat gain represents heat storage and causes an increase in node

temperature

.

The test facility is modeled using 25 nodes. While this is an

approximation of the actual facility, the results, as will be shown in

section 7, compare favorably to the measurements. Table 4.1 lists the

nodes used in the model.

Table 4.1 Nodes for Test Facility Simulation

Node

12

3

4

5

6

7

8

9

10

11

12

13

14

Location

Room Air

Plenum Air

Luminaire Air

Lamp

Ballast

Luminaire Housing

Luminaire Diffuser

Floor Surface

Floor Middle

Floor Bottom

Ceiling Bottom

Ceiling Middle

Ceiling Top

Suspended Ceiling Top

38
15

16

17

18

19

20

21

22

23

24

25

Suspended Ceiling Bottom

East Wall

South Wall

West Wall

North Wall

East Plenum Wall

South Plenum Wall

West Plenum Wall

North Plenum Wall

Plenum Steel

Luminaire Housing Top

The major simplifications related to the definition of the nodes involve

the luminaire and the room and plenum air. The four luminaires of the

test room are simulated as a single luminaire with four times the power

input, light output and area of a single luminaire. The luminaire

itself consists of six nodes, namely, ballast, housing, lamp, diffuser

top and bottom, and air. A comprehensive modeling of the luminaire

would involve detailed three dimensional conduction and convection, and

is beyond the scope of the present research effort. However, the six

node model of the luminaire proved to be useful for analyzing the

interaction of the lighting and HVAC systems, by allowing for sufficient

detail to provide sensitivity to the parameters being varied during the

testing.

The room air is assumed to be fully mixed and held at a constant

temperature. This is a widely utilized assumption, since a more

vigorous treatment would significantly complicate the problem, because

three dimensional air circulation calculations would be required. Given

the nature of the test facility being modeled, the expected increase in

simulation accuracy would be much less than the increase in model

complexity.

The plenum air is also assumed to be fully mixed, but the temperature is

allowed to vary as heat is stored or reclaimed. Again, for the purpose

of evaluating energy transfer from lighting, this assumption can be

justified.

4.2 Solution Procedure

Initially, all node temperatures are set equal to the room air

temperature, which is held constant throughout the simulation. Heat

transfer between nodes by conduction, radiation and convection is

computed for each time step. The net heat gain to each node is used to

calculate new temperatures for each node, or cooling load.

The simulation is initialized with values for nodal properties,

including surface areas, volumes, thermal capacitances, surface

emittances and radiation shape factors between nodal pairs. The

radiation shape factors were precomputed, assuming diffusely reflecting

surfaces, using the procedure from [Wa86], Also specified as input

parameters are the fractions of the visible light output from the

luminaires which are absorbed by each node. These fractions are constant

and were determined from measurement using photometers on a two foot

grid across all surfaces. The luminous efficacy of the visible light

output must also be specified. This is not the system luminous

39
efficacy, but rather the lumens contained in each watt of visible light

from the luminaire

.

All nodal temperatures are initially set equal to the room air

temperature set point. The lamp temperature determines the relative

light output, in lumens and power consumption, in watts. The form of

these temperature dependent functions is described later. The watts of

visible light output (VW) are determined by dividing the lumen output by

the luminous efficacy:

VW =

lumen output

luminous efficacy of light output (4.1)

The heat source at the lamp is determined by subtracting the watts of

visible light and the ballast losses from the lighting power. A

constant ballast fraction of 0.15 was assumed, in the absence of any

data on ballast temperature dependence.

The black body equivalent thermal radiation power (EB) from each node

(i) was computed from:

EBi = aTi4

(4.2)

where: a = Stefan Boltzmann constant

T^= absolute temperature of node i (or)

The visible light watts incident on each node (QV^) were determined

from:

QVt = SiVW (4.3)

where: = fraction of visible light absorbed by node i

(determined from measurements)

The convection coefficients between air and surface nodes are computed

from free or forced convection correlations, adjusted as appropriate

[As77]

.

Table 4.2 lists the correlation used for each node.

Table 4.2 Convection Correlations for Simulations

Lamp, Housing h = 0.18(AT)0*33

Vertical Surfaces h = 0.19(AT)^”^

Horizontal Surfaces h = 0.22(AT)^-33

AT = temperature difference , °F

The calibration of the computer model is accomplished primarily through

the adjustment of the convection coefficients, since these have the most

uncertainty and least link to fundamental heat transfer theory. The

calibration procedure is described in chapter 7.

40
The rate of heat transfer between nodes by thermal radiation is

determined by solving for the radiosity of each node. Nodal radiosity

(Bj_ ) is composed of emitted and reflected radiation, according to:

V {l>V EBJ

+JiV B

j} [ttw)

(4 - 4)

where: EF^ = EM^/(1 - EM^)

EM^ = emittance of node i

F^j = radiation shape for factor between nodes i and j

Application of equation 4.4 to all the surface nodes leads to a set of

simultaneous equations for B. These are solved using Gauss -Seidel

iteration.

The net radiation at each node (QRN^) is computed by subtracting

radiosity from irradiance, where irradiance (Hi) at node (i) is given

by:

Thus

:

N

H. =.L F. . i J-l ij

B.

3

QRNi = (Hi - Bi) Ai

(4.5)

(4.6)

where

:

Ai = surface area of node i

Heat transfer by convection is determined for each surface node using

the convection coefficients and the surface and adjacent air

temperatures

.

QCi = HCi • (Tair - Ti) At (4.7)

where: QCi = convective heat gain for node i

HCi = convection coefficient for node i

Tai r = temperature of adjacent air

Similarly, heat transfer by conduction (CKi) is computed from:

N

CK. =.£

1 J-1 K. . (T.-T.) A.

J i i (4.8)

where: Kij = conductance between nodes i and j

The heat storage in each node (QSi) is determined from:

QSi = (QRNi + QCi + QVi + QKt ) (4.9)

The lamp and ballast have additional storage amounts due to source heat,

as mentioned earlier.

41
The last step in the process is to compute new node temperature,

according to:

QS

i

new X. = T. + 77-=- (4.10)

1 1 c.

1

where: = thermal capacitance of node i = (mC

p

) i

m = mass of node i

Cp = specific heat of node i

Since room air temperature is held constant, the computed heat storage

in the room air is actually the cooling load. The plenum air heat

balance also includes heat transfer due to mass flow of the return air

through the plenum. This term is:

Plenum heat = mCpr (T - T . . ) TAO (4.11) room air plenum air

where: TAO = time increment

Once the new temperatures are computed, the entire cycle is repeated for

each time step until steady conditions are achieved, thus simulating a

step response test.

42
5. Steady State Measurement Results

Steady state measurements of lighting system performance are useful

because, as will be shown later, the lighting system may be operating at

equilibrium with its thermal environment for long periods of time

.

Thus, the lighting and HVAC system should be designed and operated so

that the lighting system is at its peak efficiency, meaning that the

lamps are between 100 - 104°F.

When the lighting system is initially switched on, the luminaire

temperature increases until equilibrium conditions are achieved. This

may take several hours. The equilibrium temperatures, and related

energy flows are dependent on the room thermal environment. For the

lighting system, two of the most important room thermal parameters are

room air temperature and a return airflow rate. The return airflow

configuration also can exert a strong influence.

The effects of room air temperature, airflow rate and return air

configuration on the thermal conditions of the lighting system and room

components were examined through a series of steady state tests. The

lighting system was left on while the various parameters were changed,

and data collected until steady state conditions were observed for each

measurement configuration. At least four hours of steady conditions

were required to ensure equilibrium had occurred.

5 . 1 Lamp Compartment Return

The effects of different room air temperatures and airflow rates will be

greater for lamp compartment return than for side slots or ceiling grill

returns, because the return air circulates directly over the lamps. The

airflow rate was varied from 120 to 160 to 200 cfm, or 30 to 40 to 50

cfm per luminaire. Room air temperature was set at 70, 75 or 80°F.

Lighting power, light output, average air and surface temperatures, and

minimum lamp temperatures were determined from measurements for each

configuration.

The effect of airflow rate and room air temperature on lighting power

consumption is shown in figure 5.1. The numbers next to the data points

are the actual room air temperatures . The trends are as would be

expected, with cooler air temperatures and higher air flow rates causing

greater power usage. It should be noted that light output also

increases along with power consumption as does luminous efficacy, and

thereby, lighting system efficiency. The difference between the

greatest lighting power, 767 watts (4.56 watts/ft^) and the least, 730

watts (4.35 watts/ft^) is five percent. By comparison, with the test

room fan off, the average minimum lamp wall temperature was 133°F, and

the lighting power consumption was 671 watts, only 87 percent of

maximum.

The variation in lighting power is tied directly to the minimum lamp

wall temperature, as shown in figure 5.2. The lighting power density is

plotted versus the average minimum lamp wall temperature. This is the

average of the individual lamp temperature minima for each luminaire

.

The range in average minimum lamp temperatures is not great, spanning

less than ten degrees F. This indicated that even the 80°F room

temperature and 120 cfm airflow condition cool the lamps to some extent,

43
with an average minimum lamp wall temperature of less than 111°F.

According to the formula from [Mu85] , a ten degree F difference in

minimum lamp temperature would correspond to a four percent change in

luminous efficacy. Least - squares linear regression was used to

correlate lighting power with minimum lamp wall temperature for use in

the computer model simulations. Relative lighting power, that is the

ratio of actual to maximum lighting power, was found to be related to

average minimum lamp wall temperature (Tj^y) by:

Relative

Lighting = 1.3089 - 0.00323 TMLy (°F) (5.1)

Power

for TVjrrTT > 95 . 6°F

MLW

Lighting system light output varied with room air temperature and

airflow rate in a manner similar to lighting power consumption, as shown

in figure 5.3. The difference between the greatest light output, and

the least is four percent. The relationship between light output and

average minimum lamp wall temperature is shown in figure 5.4. Using

linear regression relative light output was found to vary with minimum

lamp wall temperature according to:

Relative

Light = 1.2960 - 0 . 00297TMLW(°F) (5.2)

Output

for Tf£Ly> 99 . 7°F

The conditions in this series of tests found the average minimum lamp

wall temperatures to be at or above the temperature corresponding to

maximum power consumption and light output. Thus, cooling the lamps by

lowering the air temperature or increasing air flow always caused an

increase in light output and power consumption. With lower lighting

power densities, it is likely that lamp temperatures would be below the

temperatures corresponding to maximum light • output and power

consumption, under some conditions.

The relationship between lighting system luminous efficacy and average

minimum lamp wall temperature is displayed in figure 5.5. Considerable

scatter in the data is apparent, due to uncertainty in the lighting

power and light output values which make up the luminous efficacy ratio.

Due to the scatter and limited data available, no correlation was

derived for luminous efficacy. However, it appears that luminous

efficacy peaks near 104°F, with a minimum value, for the lowest average

lamp wall temperature (~98°F) , two percent lower than the maximum. This

low lamp temperature occurred for an extra data point collected at a

room temperature of 65°F. A decrease in luminous efficacy is also seen

at the hotter lamp temperatures.

Average minimum lamp wall temperature as a function of airflow rate and

room air temperature is shown in figure 5.6. A five degree change in

room air temperature causes a correspondingly smaller change in minimum

lamp temperature , about four degrees at the high airflow rate , and three

degrees at the low airflow rate. Changing airflow rate has the greatest

effect at the 70°F air temperature, with an increase in airflow from 30

44
to 50 cfm per luminaire causing a 3°F drop in minimum lamp temperature.

The same airflow change caused only a 1°F drop in minimum lamp

temperature for the 80°F room air temperature. This is probably due to

the stronger effects of convection when the room and lamp temperature

differences are greater.

While the room air is being held constant at the set - point, the plenum

air can heat up due to energy from the lights. Figure 5.7 shows average

plenum air temperature, computed from the average of 16 thermocouples,

as a function of airflow rate and room air temperature. At the lowest

airflow rate, plenum air temperature increased by between 5 and 6

degrees F over room air temperature. At the highest airflow rate,

plenum air temperature exceeded room air temperature by 4 to 5 degrees

F.

Figure 5.8 shows the average floor temperatures for both the tops of the

floor slab and ceiling slab. Very close agreement is seen between the

average temperatures of the top surfaces of both slabs. Average floor

surface temperature is about two to three degrees F warmer than room air

temperature, with the extreme range in airflow rates causing less than a

one degree change in floor temperature , for a constant room air

temperature

.

Similar results are shown for the ceiling in figure 5.9, which shows

data for both the bottoms of the ceiling and floor slabs. The match in

temperatures is not as good as the floor temperatures . This is partly

due to the high conductivity steel decks, which make surface temperature

measurements difficult, partly due to the small number of thermocouples

under the floor slab, and partly due to control system performance. The

lower steel deck is typically measuring one degree warmer than the upper

steel deck. The upper steel deck (room ceiling) is about two degrees

above room air temperature.

The average temperature of all the walls was determined individually and

as a group, both for above and below the suspended ceiling. Figure 5.10

shows the results for the average temperature of all of the walls

together. Below the plenum, the room walls do not show much increase in

temperature due to operation of the lighting system. The walls in the

plenum do heat up noticeably.

Some temperatures are at or slightly below the room air temperature setpoint. It is possible that the wall temperatures could be slightly

below the average room air temperature due to non -uniform room air

temperatures. The walls above the plenum heat up about two or three

degrees above room air temperature. The full range of airflow rates

have about a one degree effect, with the highest airflow rate keeping

the plenum air about one degree cooler than the lowest airflow rate.

Average temperatures were also determined for each wall individually.

The results are shown for the north wall in figure 5.11, east wall in

figure 5.12, south wall in figure 5.13 and west wall in figure 5.14.

The west wall is not instrumented with as many thermocouples as the east

wall, and the north and south walls have detailed thermocouple grids

over only one half of their widths. The individual wall temperatures

are similar to the average wall temperatures, although the north and

east walls are shown to be slightly warmer than the south and west

45
walls. In all cases the plenum walls are about three degrees F warmer

than the corresponding room walls. Also, a five degree change in room

air temperature produces nearly a five degree change in average wall

temperatures, for all conditions.

5 . 2 Effect of Return Configuration

A series of steady state tests was run with different airflow rates and

return air configurations, for a room temperature of 75°F, with the

lights on. Figure 5.15 presents lighting power as a function of airflow

rate and return path. The letters next to the points identify the

return air path by its initials . Lighting power with the lamp

compartment return is substantially greater than either the side slot or

ceiling grill return. This is due to the increased cooling of the lamps

with the lamp compartment return. The lighting power results are very

similar for the side slots and ceiling grill returns. Lighting power

increases with airflow for all return configurations.

The variations in lighting power are a direct result of the minimum lamp

temperatures, as shown in figure 5.16. Minimum lamp wall temperature

decreases with increasing airflow rates, with lamp compartment return

keeping the lamps about 7 oF cooler than either the side slots or

ceiling grill. The range of airflow rates caused minimum lamp wall

temperature to vary by 3 °F for each return configuration.

Figure 5.17 presents average plenum air temperature versus airflow rate

and return path. Plenum air temperatures decrease with higher airflow

rates, particularly for the lamp compartment return. Plenum air

temperatures span about 2 °F for the range of conditions, and are

greatest for the lamp compartment return. This is due to the heat

pickup from the luminaires. Plenum air temperatures for the ceiling

grill return are affected the least by airflow rate.

The ceiling and floor temperatures are shown in figures 5.18 and 5.19,

respectively, versus airflow rate and path. The scales are magnified on

these plots. Both the floor and ceiling heat up the most for the

ceiling grill return at 200 cfm, and the lamp compartment return at 120

cfm. These two opposite conditions have similar effects because low

airflow with the lamp compartment return allows the slabs to heat up,

while the high airflow with the ceiling grill increases convective heat

transfer to the slabs from the warm plenum air. The slab temperatures

exceed room air temperature by two to four degrees F for the range of

conditions tested.

Figures 5.20, 5.21, 5.22 and 5.23 present the temperature results for

the north wall, east wall, south wall and west wall, respectively, both

for the room and plenum portions of the wall. The patterns are very

similar for each of the walls, with the plenum walls being about 3 °F

warmer than the room walls. The low temperatures indicated for the

south and west walls are due to the limited number of temperature

sensors located on these surfaces. The return air path has very little

effect on the wall temperatures. Increasing the airflow rate causes

plenum wall temperatures to decrease, but has little effect on room wall

temperatures. This indicates that heat transfer by forced convection at

the room walls is a minor effect.

46
5 . 3 Luminaire Temperatures

Luminaire temperatures were measured using thermocouples distributed on

the luminaire surfaces. As was mentioned earlier, luminaire

temperatures reach equilibrium relatively quickly after the lights are

switched on or off. Figure 5.24 shows the rapid rise in luminaire

temperature after the lights are switched on. Room air temperature is

70°F, and airflow equal to 200 cfm with ceiling grill return for this

test. The two warmest temperatures are the tops of the lamps, which are

above 52°C or 126°F. The bottom of the lamp end is 46.5°C, or 116°F.

At the center of one lamp, the top temperature is 52.4°C (126°F) and the

bottom 43 . 5°C (109°F). For the second lamp, temperatures at the same

locations are 43 . 2°C (110°F) and 37.0°C (99°F) . This demonstrates the

cooling effect at the bottom of the lamps, and the trapping of heat near

the lamp tops. The top of the luminaire housing, facing the plenum, is

at 39°C, or 102F, and the diffuser is nearly at room air temperature,

21 . 6°C or 71°F.

Similar data are presented in figure 5.25 for side slot air return, 200

cfm and 70°F test condition in a lights off test. The two warmest

temperatures are again the lamp tops, at 52°C or 126°F. The ballast

housing interface is 47.7°C or 118°F. The bottom of the lamp center is

43°C (110°F), while the housing bottom is approximately 36°C (97°F)

.

All of the temperatures return to room temperature in less than two

hours, except for the ballast which requires about four hours.

Similar luminaire temperature distributions were observed for other test

configurations. Fluorescent lamp temperatures varied substantially from

top to bottom, and end to center, and from lamp to lamp. These

temperature variations are due to localized differences in airflow

patterns and rates. As a result, the entire lighting system, including

16 lamps, is operating with each lamp drawing a different amount of

power as a result of its individual minimum lamp wall temperature. The

variations in lighting power and light output observed were for the

lighting system as an ensemble.

47
WATTS

Figure 5.1 Lighting Power Versus Room Air Temperature and Airflow Rate

for Lamp Compartment Return

Figure

Lighting power

100 120 140 160 160 200 220 MEASURED MR FLOW, CFM

REL LIGHT POWER VS AVG MIN LAMP TEMP

LAMP COMP RETURN

5.2 Lighting Power Versus Average Minimum Lamp Wall Temperature

48
LUMENS/MAX

LUMENS

LUMENS/

SQM

Figure 5 .

3

Light Output Versus Room Air Temperature and Airflow Rate

LIGHT OUTPUT VS AIRFLOW AND RM TEMP

96.5

96

97.5

97

96.8

96

96.6

98

94.8 100 120 140 160 180 200 220

AIRFLOW RATE, CFM

70 F T 76 F O 80 F

1.01

1.008

1

0.995

0.99

0.988

0.98

a978

0.97

0.966

096

Figure 5.4 Light Output Versus Average Minimum Lamp Wall Temperature

REL LIGHT OUTPUT VS AVG MIN LAMP TEMP

LAMP COMP RETURN

LAMP COMP RETURN

49
RELATIVE

LUMINOUS

CrFICACY

Figure 5.5 Lighting System Luminous Efficacy Versus Average Minimum

Lamp Wail Temperature

REL LUMINOUS EFFICACY VS MIN LMP TEMP

LAMP COMP RETURN

1.024

1.022

1 02

1.01 a

1.016

1.014

1.012

1.01

1.006

1.006

1.004

1.002

1 97 99 1 01 103 '106 107 109 111

AVG MM LAMP WALL TEMP. DEO F

MINIMUM LAMP TEMP.

OA-CH, LAMP COMPARTMENT

111

110

109

108

107

106 6

hiO 105

104

103

102

101

100 100 120 140 160 180 200 220 MEASURED AIR FLOW. CFM

Figure 5.6 Average Minimum Lamp Wall Temperature Versus Room Air

Temperature and Airflow Rate

50
OEG.

F

QEG.

Figure 5 .

7

Average Plenum Air Temperature with Lights on Versus Room

Air Temperature and Airflow Rate

AVERAGE PLENUM AIR TEMP.

DA— 0(. LAMP COMPARTMENT

Average floor temperature.

DA- CX, LAMP COMPARTMENT

MEASURED AIR FLOW, CFM Q ASOVE PLENUM + BELOW PLENUM

Figure 5 . 8 Average Floor Temperature with Lights on Versus Room Air

Temperature and Airflow Rate

51
oeo.

F

OEG.

Figure 5,9 Average Ceiling Temperature with Lights on Versus Room Air

Temperature and Airflow Rate

Average ceiling temperature

DA—Ot. LAMP COMPARTMENT

MEASURED AIR PLOW, CFM O ABOVE PLENUM -t- BELOW PLENUM

Average wall temperature

DA- CM, LAMP COMPARTMENT

MEASURED MR FLOW. CFM

ABOVE PLENUM +• BELOW PLENUM

Figure 5.10 Average Wall Temperature with Lights on Versus Room Air

Temperature and Airflow Rate

52
DEG.

F

Ota.

Figure 5.11

ik

Figure

Average North Wall Temperature with Lights on Versus Room

Air Temperature and Airflow Rate

North Wall temp.

OA-Oi LAMP COMPARTMENT

09

84

&3

82

8180

797877

76

75

74

7372

7170

100 120 140 160 180 200 220 240 MEASURED AIR FLOW, CFU

ABOVE PLENUM + BELOW PLENUM

800

79.99 0

800

+•80 + 79.99 + 80

750

750

76 0

71 -36 O

+•75 + 75 + 75

69.99 0

70

+ 71.36

+ 69.99 + 70

T —I 1 1 1 T

1 r— 1— 1

' I r

East Wall temp.

OA—01, LAMP COMPARTMENT

MEASURED AIR FLOW. CFM

Cl ABOVE PLENUM + BELOW PLENUM

.12 Average East Wall Temperature with Lights on Versus Room

Air Temperature and Airflow Rate

53
OEG

f

OEO.

Figure 5.13 Average South Wall Temperature with Lights on Versus Room

Air Temperature and Airflow Rate

South Wall temp.

DA— 01, LAMP COMPARTMENT

MEASURED AJR FLOW, CFM

ABOVE PLENUM SLOW PLENUM

West Wall temp.

DA— D», LAMP COMPARTMENT

MEASURED AIR FLOW, CFM

ABOVE PLENUM + BELOW PLENUM

Figure 5.14 Average West Wall Temperature with Lights on Versus Room

Air Temperature and Airflow Rate

54
WATTS

LIGHTING POWER VS AIRFLOW RATE AND PATH

LAMP COUP. SO£ SLOTS OR COL CRUL

LMP CUP

AIRFLOW RATE. CPU

S SLOTS CEL GRL

Figure 5.15 Lighting Power Versus Airflow Rate and Path

55
Figure 5.16 Average Minimum Lamp Temperature Versus Airflow Rate and

Path

t

tf

£

Min. lamp temp.

STEADY STATE

f

r !

St.fi Plenum air temp.

tmc* STATE

St -

SO.fi8

ue

8

1£

7» -

ms -

o

co

LC

o

o

00

i > I I T~~T 1 | ' I T—Tf—

vm m t«o t«o ise aoo sao s«s

Figure 5.17 Plenum Air Temperature Versus Airflow Rate and Path

56
oca.

r

Figure 5.18 Ceiling Temperatures Versus Airflow Rate and Path

Ceiling surface temp.

steady state

Floor surface temp.

STTOT STATE

a MOvcniMUH + sojov^enum

Figure 5.19 Floor Temperatures Versus Airflow Rate and Path

57
Figure 5.20 North Wall Temperature Versus Airflow Rate and Path

North Wall surface temp.

STEADY STATE

Mr «#m

a MM PUENUM BELOW SLEMUM

East Wall surface temp.

STEADY STATE

Figure 5.21 East Wall Temperature Versus Airflow Rate and Path

58
Figure 5.22 South Wall Temperatures Versus Airflow Rate and Path

South Wall surface temp.

. STEADY STATE

West Wall surface temp.

STEftnr STATE

Figure 5.23 West Wall Temperatures Versus Airflow Rate and Path

59
C

Figure 5.24 Luminaire Temperatures During Lights on Step Change Test

u?

to — @2te — ©3te mM® —

1 06te 06ta

cj

twwp* (9E) SO

66

SO

4S

4©

36

30

26

20 @ 2 « « • 10 <2 14 16

tima.hrB mm " Ifwp «ng " Imp «*4 ^

Figure 5.25 Luminaire Temperatures During Lights off Step Change Test

60
6. Transient Measurements

Transient tests were run to determine the response of cooling load to a

step change in lighting. Starting from a steady state condition with

lights on or off, the lighting was switched and the cooling load

monitored until equilibrium conditions were attained. Usually, back-toback tests were run, first a lights-on test, followed by a lights-off

test. The lighting system is the only planned heat source in the test

room. Under ideal conditions, the measured cooling load with the test

room at equilibrium and the lights off would be zero. At equilibrium

with the lights on, cooling load would equal lighting power. In

reality, small heat gains to the test room caused slight cooling loads

with the lights off. This zero offset was assumed to be constant for

the duration of a test, and the measured cooling loads were adjusted

accordingly. The typical zero offset was less than 20 watts, compared

to a lighting power of 750 watts.

After adjusting for the zero offset, the measured cooling load at

equilibrium with the lights on did not precisely match the measured

lighting power. This offset was assumed to be due to variations in the

conversion factor used to compute return airflow volume from the

measured average velocity and heat loss induced by the lights-on

condition. The measured cooling loads were adjusted so cooling load

equaled measured lighting power with the lights on and the test room at

equilibrium. This correction factor was usually less than four percent,

and varied with airflow rate, increasing slightly as airflow increased.

Figure 6.1 shows the measured cooling load and lighting power for a

lights-on test followed by a lights-off test. The lighting power

follows the square wave pattern, and the resulting cooling load an

exponential pattern. The spikes in the cooling load before the lights

were turned on occurred during routine maintenance and are of no

consequence to the test. Slight cooling loads are seen when the lights

are off, both before the initial switching of the lights and after the

lights were turned off. Measured cooling load was less than measured

lighting power with the lights on. After adjusting for the zero- and

full scale offsets, the cooling load appears as in figure 6.2. This

adjustment procedure was followed for all of the test results.

The figure labels sometimes include a two character code, such as EM.

This is simply a test identification code, and is not important. The

label PARA refers to parabolic diffuser type luminaire. The code in the

figure captions relates to the airflow rate and return configuration,

according to:

CG ceiling grill return

SS side slot return

LC lamp compartment return

120 - 120 cfm airflow

160 - 160 cfm airflow

200 - 200 cfm airflow

6.1 Cooling Load Profiles and Regression Analysis

Preliminary analysis of the initial transient tests indicated that a

single exponential relation was not sufficient to represent cooling load

61
due to lighting during the first several hours after the lighting system

had been switched on or off. This fact was noted by other researchers,

as mentioned earlier. This effect is demonstrated in figure 6.3, which

shows measured cooling load and lighting power and calculated cooling

load using a single exponential, as in equation 2.1. The test

conditions are ceiling grill return, room air temperature of 75°F and

return airflow rate of 160 cfm. The calculated cooling load

overestimates the measured cooling load for the first two or three

hours. The regression coefficients for this test were A = 0.46464 and B = 0.16534.

The difference between ‘the calculated and measured cooling loads is

plotted in figure 6.4. This initial transient is due to heat storage in

the luminaires and heat absorption by. the light-weight room components,

such as the walls and suspended ceiling. The shape of this plot

suggested the addition of a second exponential relation to match the

initial transient response. Thus, cooling load due to lighting would be

given by:

= 1 - Ae“

Bt - Ce“

Dt

(6.1)

w

The values of C and D can be determined from linear regression,

according to:

In [1 - Ae”

Bt

- Q(t) ] = In C - Dt (6.2)

W

Values for C and D were computed for this test, and the resulting

calculated cooling load plotted versus measured cooling load in figure

6.5. The agreement between measured and calculated cooling loads is

quite good. The regression coefficients were C = 0.41984 and D = 1.21765.

Before continuing with the results from other transient test, the

derivation of the weighting factors corresponding to equation 6.1 is

warranted. Equation 6.1 is the output (cooling load = Q(t)) of the

system to a step lighting input (W)

.

In the discrete case:

Q(z) G(z)

z-1

z z-1

Az Cz

-BT -DT

z-e z-e

(6.3)

The transfer function G(z) is needed to determine the weighting factors,

so that the response of cooling load to arbitrary lighting power input

patterns can be computed. Thus,:

G(z) = 1

Az ( z - 1

)

Cz (z-1

)

, -BT. , -DT,

z (z-e ) z(z-e )

, 2 -BT W 2 -DT. /a 2 . w 2 -DT. 2 . w 2

(z -ze ) (z -ze ) - (Az -Az)(z -ze )-(Cz -Cz)(z -ze ___)

(6.4)

-BT.

, 2 -BT.

.

2 -DT.

(z -ze )(z -ze ) (6.5)

62
Simplifying and combining terms, we obtain:

o , -DT -BT „ -DT -BT -1 , - (B+D)T . -DT _ -BT. -2

x ( l-A-C)z +(Ae +A+Ce +C-e -e )z + (e -Ae -Ce )z

G(z)_

-DT. -BT V -1 . , -(B+D)T. -2

° -1

a z +a.z +a„z

o 1 2

1 - b^z *+ b

2 z

^

1 - (e + e )z + (e

-2

)z ( 6 . 6 )

(6.7)

where

:

aQ = (1 - A - C)

a? = (Ae*DT+ A + Ce-*T+ C - e

- DT - e*“)

ao = (e-(B+D) T - Ae-DT. Ce'ET)

bT = (e- DT+ e-BT)

b2 = (e-(B+D)T)

Inverting back to the time domain, the cooling load due to lighting is

given by:

Q(t) = a0W(t) + aiW(t-T) + a2W(t-2T) + bxQ(t-T) - b 2Q(t-2T) (6.8)

The quantities aQ , a^ , a2 , b^ , and b 2 are the weighting factors

corresponding to equation 6.1. The addition of the second exponential

introduces two additional terms, namely the lighting power and cooling

load from two hours before the current time step. The five weighting

factors are not all independent, since they are derived from only four

coefficients. The values of the weighting factors are determined,

however, strictly from the regression coefficients A, B, C and D, and

choice of a time step increment. A one-hour time increment is

frequently employed for building energy analysis, although any time step

can be used.

The values for A and B are, of course, the same for the single and

double exponential relations for computing cooling load due to lighting.

Thus, after the initial transient has dissipated, the cooling loads due

to a step change in lighting are equivalent. The traditional ASHRAEtype weighting factors can be computed from A and B.

Returning to the transient test results, a lights-off test was run with

identical conditions to the first test described above. The two tests

were run back-to-back. Figure 6.6 displays measured cooling load and

lighting power, and cooling load calculated by a single exponential

regression fit. The regression coefficients were A = 0.62233 and B = 0.20377. Cooling load is underestimated for several hours following the

turning off of the lights. In the case of the double exponential fit

for the lights-off test, cooling load is given by:

Q(t) = Ae“ BT + Ce" DT (6.8)

The regression coefficients can be determined as before. After

subtracting the first exponential, the initial transient was plotted, as

in figure 6.7. Performing a second regression analysis, the value of C

was found to be 0.38232 and D was found to be 2.24065. The double

exponential regression fit is plotted in figure 6.8. The initial

transient decayed slightly faster during the lights-off test. However,

measurement uncertainty and non-perfect control system performance,

63
however good, combine to obscure the significance of this difference.

Although the regression coefficients from the two tests are not

identical, the cooling load profiles they represent are very similar, as

will be shown later.

Similar tests were run with ceiling grill return, but with various

airflow rates. Figure 6.9 displays measured and calculated cooling

loads for a lights-on test at 120 cfm, for the single exponential. The

initial transient is of a very short duration. The double exponential

fit is shown in figure 6.10. The value of D = 4.250 was the largest for

all the transient test runs. The values of the other coefficients were

A = 0.72791, B = 0.24514 and C = 0.14638. For the ceiling grill return,

200 cfm test, the measured and single exponential calculated cooling

loads are plotted in figure 6.11. The regression coefficients are A . =

0.56150, and B = 0.18567. The initial transient is shown in figure

6.12, and the double exponential fit is plotted in figure 6.13. The

agreement between the measured and calculated cooling , loads is very

good. The value of C was 0.39523 and D was 2.04771.

A similar series of tests was run with side slot and lamp compartment

returns, for the same three airflow rates. The regression coefficients

are summarized following the description of the figures. Figure 6.14

shows the measured cooling load and single exponential fit for the

lights-on the test with side slot return at 120 cfm. Figure 6.15 shows

the initial transient, and 6.16 displays the double exponential fit.

For the same condition, lights -off test, the measured and single

exponential cooling loads are plotted in figure 6.17. The initial

transient is shown in figure 6.18, and the double exponential fit in

figure 6.19.

For this test, and almost all of the others, the initial transient

decays faster during the lights-on test than the lights-off test. This

is partly due to the sharp peak in lighting power when the lights are

initially energized. No equivalent "negative peak" occurs when the

lights are extinguished. It is also possible that some physical

mechanism inhibits heat storage when the lights are initially turned on,

and retards the discharge of stored heat when the lights are turned off.

A third possibility would be some artifact of the control system. A

precise pinpointing of the cause of this difference was not forthcoming,

however, the impact on the average cooling load profiles was small, as

will be demonstrated later.

Figure 6.20 shows the measured cooling load and single exponential fit

for the side slot return at 160 cfm, lights-on test. Figure 6.21

displays the initial transient, and 6.22 plots the double exponential

fit. The same three plots for the lights-off test are shown in figures

6.23, 6.24 and 6.25.

The results for the side slot tests with 200 cfm airflow are shown is

figures 6.26, 6.27 and 6.28 for the lights-on test and 6.29, 6.30 and

6.31 for the lights-off test.

Similar data for the lamp compartment return are displayed in figures

6.32, 6.33 and 6.34 for 120 cfm, lights- off test. The test had the

poorest match between the measured cooling loads and the double

64
exponential fit. The cooling load profile is characterized by a rapid

initial change, followed by a prolonged, slow settling to zero.

Figures 6.35, 6.36 and 6.35 present the data for the 160 cfm, lights -on

test, and figures 6.38, 6.39 and 6.40 show the results from the 160 cfm,

lights-off test. Finally, figures 6.41, 6.42 and 6.43 display the

results for the 200 cfm, lights-off test. In all cases, the agreement

between the measured cooling load and double exponential fit was very

good.

Since, in most cases, more than one test was run for each configuration,

suitable average regression coefficients needed to be determined to

represent each condition. Simple arithmetic averaging of multiple

values of each regression coefficient was found to be sufficient for

this purpose. A demonstration of this also addresses the issue of the

significance of different regression coefficient values from different

tests with identical configurations, and explores the repeatability of

the measurements

.

A series of four tests were run in succession with identical test

configuration of side slot return and 200 cfm airflow. Two of the tests

were lights -on and two were lights-off. The test identification codes

and regression coefficients are listed in Table 6.1.

Table 6.1 Regression Coefficients for Side Slot

200 cfm Tests

ID Code

ES On 0

ES Off 0

ET On 0

ET Off 0

A B

58987 0.21513

64542 0.23799

642139 0.23166

57479 0.21544

C

0.31755

0.35654

0.25832

0.3558

D

1.76135

2.70828

2.69113

1.88319

While the regression coefficients vary for the four tests, the cooling

load profiles they represent are very similar. Figure 6.44 shows the

cooling loads for each of the four tests, plotted as symbols, and the

cooling load plotted from the average regression coefficients, plotted

as a solid line. Very good agreement is apparent between the tests, and

the cooling load profile calculated from the average coefficients

represents the average of all the cooling load profiles. This figure

indicates good measurement repeatability.

Average regression coefficients were determined for each test

configuration. These values are summarized in table 6.2, along with

overall average values for each regression coefficient.

65
Table 6.2 Average Regression Coefficients

for Each Test Condition

Return Airflow A B C D

CG 120 0.72791 0.24514 0.14638 4.250

CG 160 .53937 .18183 .39826 1.71436

CG 200 .57391 . 20444 .45193 2.47985

SS 120 .68507 .21748 .37771 3.1224

SS 160 .68603 .22321 .40393 4.010

SS 200 .61305 .22506 .32205 2.2610

LC 120 .47603 .11463 .52696 1.21763

LC 160 .60498 .18701 . 34944 1.9729

LC 200 .63085 .22480 .34122 2.93645

AVG 0.615311 0.202622 0.368653 2.662736

CG - Ceiling Grill/SS - Side Slots/LC - Lamp Compartment

The effect of the airflow rate and return air configuration on the

regression coefficients was examined in a series of plots. Figure 6.45

shows all the regression coefficients plotted versus airflow rate. All

return air paths are included in this figure. The variability and

magnitude of the D coefficients tends to obscure the differences between

the other coefficients. The variability of the D values is least for

the 200 cfm airflow. Figure 6.46 displays the values of A, B and C for

all tests. In general, these coefficients bunch together at the highest

airflow rate also.

The effect of airflow rate and return air path on the individual

regression coefficients is shown in figures 6.47 for A, 6.48 for B, 6.59

for C and 6.50 for D. No clear pattern is apparent relating the values

of these coefficients to airflow rate. The A values tend to increase

with airflow rate for the lamp compartment return, and decrease for the

side slots and ceiling grill returns. The B values vary the least. The

value of B is dependent on the long - term heat storage characteristics

of the room, which are only weakly influenced by airflow rate of the

path. The value of C increases with airflow rate for the ceiling grill

return, and decreases for lamp compartment return. The opposite is true

for the D values

.

In order to compare the cooling load profiles represented by the

regression coefficients, cooling load responses to a unit step change in

lighting were computed and plotted for each test configuration. Figure

6.51 shows the cooling load profiles for the ceiling grill return,

figure 6.52 for the side slots and figure 6.53 for the lamp compartment

return. The effect of airflow rate is minor, except for the lamp

compartment return. The cooling load profile for the 120 cfm airflow

rate is consistently less than the other airflow rates, but not by a

large margin.

A second series of plots compares the effects of return air path for

constant airflow rate. Figure 6.54 is for 120 cfm, 6.55 for 160 cfm and

6.56 for 200 cfm. Only the lamp compartment return, 120 cfm cooling

66
load profile stands out, being noticeably lower than the side slots or

ceiling grill return.

One general indicator of short term cooling load performance is to look

at the cooling load for one hour after the lights are energized. Figure

6.57 shows cooling load after one hour versus airflow rate and return

air path. The effect of airflow rate is quite strong and apparent, with

increasing airflow rate causing a greater cooling load after one hour.

The difference between the extreme values is approximately 20 percent.

The side slot return has the lowest cooling load at each airflow rate.

The ceiling grill return produces the greatest one hour cooling load at

120 and 160 cfm, while the lamp compartment is greatest at 200 cfm.

6.2 Weighting Factors and Cooling Load Fractions

Weighting factors were computed for each of the test configurations

using equation 6.7, based on the average values for the regression

coefficients. These weighting factor values are listed in table 6.3.

Table 6.3 Average Weighting Factors

For Each Test Condition

ao a l a2 b l b 2

CG120

CG160

CG200

0.12571

0.06237

-0.02584

0.20237

0.35298

0.54342

-0.11377

-0.27903

-0.348L6

0.796859

1.013821

0.898859

0.011163

0.150139

0.068269

SS120

SS160

SS200

-0.06338

-0.08996

0.0649

0.54887

0.60744

0.35344

-0.29864

-0.32105

-0.23781

0.848593

0.818080

0.902714

0.035439

0.014505

0.083237

LC120

LC160

LC200

0.00299

0.04558

0.02793

0.42612

0.35989

0.42633

-0.34687

-0.25862

-0.26362

1.187626

0.968488

0.851729

0.263880

0.115335

0.042372

AVG 0.016035 0.44158 -0.28699 0.886343 0.056962

CG-Ceiling Grill/SS-Side Slots/LC-Lamp Compartment

The effect of airflow rate and return air configuration on the weighting

factor values was examined in a series of plots. Figure 6.58 shows a0

versus airflow for all return air paths. The aQ values are small, and

bunch together at the 200 cfm airflow rate. The three negative values

are not consistent with the physical system, but result from the

extremely steep initial slope of the cooling load. Based on observation

of these aQ values, the substitution of an a2 term for an aQ term

utilized in [Mi74] does not seem unreasonable.

The values of are shown in figure 6.59. Again, the values converge

somewhat at 200 cfm. The values of a2 , b^ and b 2 are displayed in

figures 6.60, 6.61 and 6.62, respectively. No general correlations are

apparent, although the trend for values to converge at the 200 cfm

airflow rate continues to be seen. The b^ and b 2 values exhibit a

similar pattern, due to their interdependence.

67
In order to evaluate the effect of the airflow rate and configuration on

cooling load due to cyclic operation of the lights, cooling load

fractions were computed from the weighting factors for a ten hour on

daily lighting cycle. Figure 6.63 shows the hourly cooling loadfractions, that is the ratio of cooling load to lighting power, for four

successive daily cycles. The data repeat after the second day. This

plot is for one test configuration. What is of interest is the steady

periodic cooling load, which occurs in this plot from the second day on.

Peak cooling load fraction does not reach 1.0, and some cooling load is

present at all times, even when the lights are off.

Figures 6.64, 6.65 and 6.66 show the steady periodic cooling load

fractions versus airflow rate for ceiling grill, side slots and lamp

compartment, respectively. The most notable aspects of the cooling load

fraction profiles is their similarity. With the exception of a few

points at the lights on and lights off instants, and the 120 cfm, lamp

compartment return configuration, the cooling load fraction profiles are

very similar. The exceptional points occur due to the negative aQ

values. This issue is addressed below.

Figure 6.67 shows the cooling load fraction profile for the average set

of weighting factors. This is the average profile for all test

configurations

.

While the previous cooling load fractions were based on unit lighting

power, the actual cooling load profiles will vary due to differences in

lighting power for each test configuration. To examine this effect, the

cooling load fractions for each configuration were scaled by the average

measured lighting power. These results are shown in figures 6.68, 6.69

and 6.70, for ceiling grill return, side slots and lamp compartment

return, respectively.

The scaling by actual lighting power has a slight (approximately 5

percent maximum) effect which is difficult to discern from the figure.

However, it is important that this factor be considered when estimating

peak cooling loads, since the cooling load fractions only determine the

shape of the cooling load profiles, not the magnitudes. The lighting

powers and peak cooling loads for each test configuration are listed in

table 6.4.

Table 6.4 Lighting Power and Peak Cooling Load

Return Airflow (cfm)

Lighting

Power (watts)

Peak

Cooling (watts)

Ceiling Grill 120 726.5 670.0

Ceiling Grill 160 732.6 660.8

Ceiling Grill 200 734.9 671.3

Side Slots 120 728.4 660.9

Side Slots 160 730.3 665.8

Side Slots 200 733.3 676.3

Lamp Corap

.

120 753.2 686.3

Lamp Comp

.

160 755.9 676.2

Lamp Comp

.

200 762.5 701.3

68
Peak cooling load varies over six percent for the range of test

conditions, primarily due to variations in lighting power. The maximum

peak cooling load occurred for the 200 cfm, lamp compartment return,

which had the greatest lighting power. Of course, light output would be

high for this condition also.

In order to alleviate the problems due to negative a0 values, the

offending values were replaced with zeros, and the other coefficients

adjusted to compensate. The aQ values were close to zero anyway.

Compensation with and a2 were both tried. The sum of the "a"

coefficients must remain the same after adjustments. Figure 6.71 shows

the original cooling load fraction profile, and two- new profiles

corresponding to compensating with a^ and a2 , for the ceiling grill, 200

cfm test condition. The a^ compensation is seen to alleviate the

problem points while matching the original profile test. The adjusted

profiles for side slots 120 cfm and 160 cfm are shown in figures 6.72

and 6.73, respectively. While the new profiles are slightly different

than the originals, the elimination of the problem points outweighs this

aspect.

A better way to alleviate the problem of negative aQ values would be to

force the cooling load regression fit through zero. This would

correspond to having A plus C equal to one, thus aQ would equal zero.

This approach will be investigated for future test results.

ASHRAE - type weighting factors were calculated from the regression

coefficients in table 6.2, using equation 2.11. Only the values of A

and B are required for this calculation. Table 6.5 lists the ASHRAE

type weighting factors for each of the test conditions.

Table 6.5 ASHRAE - Type Weighting Factors

Return Airflow a l a2 b l

CG 120 0.430 -0.213 0.783

CG 160 0.550 -0.384 0.834

CG 200 0.532 -0.347 0.815

SS 120 0.449 -0.254 0.805

SS 160 0.451 -0.251 0.800

SS 200 0.510 -0.308 0.798

LC 120 0.576 -0.468 0.892

LC 160 0.498 -0.327 0.829

LC 200 0.496 -0.295 0.799

Average 0.498 -0.315 0.817

CG-Ceiling Grill/SS-Side Slots/LC-Lamp Compartment

The hi values cluster above 0.8, similar to the lightweight room values

from [Mi74] . Increasing airflow rate causes a^ to increase for the

ceiling grill and side slot configurations, but to decrease for the lamp

compartment return. This may be attributed to the relatively stagnant

air conditions in the test room with the lamp compartment return at the

low airflow rate. The initial heat pickup is greater, but the progress

toward equilibrium is slower, as the heat from the lights is absorbed by

the massive elements of the test room. The a^ values are similar to

[Mi74]

.

69
6.3 Heat Storage in Room Components

The heat stored in the test room components during a step response test

can be determined in several ways. The total heat stored can be

calculated from the difference between lighting power and cooling load.

The heat stored in each component can be measured using heat flow

sensors, or calculated from the measured temperature rise of each

component. While the test facility is instrumented with heat flow

sensors, there are not enough sensors to enable to measurement of all

the heat storage, thus, the first and third methods were utilized. The

heat flow sensors are useful, however, for giving a qualitative

indication of heat storage processes, at the attached surface.

The total heat stored in a lights-on test is:

Total Heat Stored = + Ce dt (6.9)

( 6 . 10 ) The heat reclaimed in a lights - off test is also given by equation

6.10. The units of the heat storage determined in this manner are

energy per unit power. In this case, that is watt-hours per watt of

lighting power. Figure 6.74 shows total heat stored versus airflow rate

and return configuration. Airflow rate has a strong effect on heat

stored with the lamp compartment return, with lower airflows causing

greater heat storage. Airflow rate has little effect on total heat

storage with side slots or ceiling grill. These effects are probably

due to the shifting of heat storage from the room to the plenum as

airflow increases. There is a slight decrease in total heat stored with

decreasing airflow rate for the side slot return, but a slight increase

for the ceiling grill return.

The heat flow sensor readings provide a view of the dynamic heat flow

into the room surfaces. Figure 6.75 shows the heat flow into the floor

and the north, east and west walls for back-to-back lights on and lights

off tests. The slight negative readings at the beginning of the tests

correspond to heat flow from the surface to the room air. These heat

flow sensors are at the center of each of the respective surfaces. Some

localized heat flow may occur, as the control thermopile is attempting

to prevent any net heat flow through the wall as a whole.

Once the lights are turned on, heat flows into all of the surfaces at a

decreasing rate until equilibrium conditions are attained. The heat

flow into the walls is much less than into the floor slab, and the walls

reach equilibrium within several hours. The initial heat flow into the

floor is very large, over 7 watts per square meter, and over 20 hours

are required for stable floor heat flow. Heat is initially stored in

the floor, from the test room and the lower plenum, until the floor

temperatures increases and heat begins to flow back into the room,

eventually at a steady rate. This net heat gain through the floor is

matched by a heat loss through the ceiling. When the lights are turned

off, heat flows back into the room as the floor cools. The amount of

heat dissipated from the floor to the room is greater than the amount

stored, since the lower plenum supplies heat to the floor also. Again,

conservation of energy is maintained by an identical loss through the

ceiling slab.

70
Heat storage in the individual room components was determined from the

measured average temperature rise between the lights-off and lights -on

condition, times the component mass and specific heat. The average of

the floor and ceiling slab temperature rises was used to compute the

slab storage, which actually includes heat stored in the ceiling and

floor slabs from the test room, but not the lower plenum or upper guard

air space. The plenum and room portions of each wall are listed

separately, since they attain different temperatures.

Table 6.6 presents the heat storage fractions in the room components for

all the test conditions. These are defined as the fraction of the total

heat stored in all the room components which is stored in an individual

component, such as the walls or slab.

Table 6.6 Heat Storage Fractions in Room Components

Ceiling Grill Side Slots Lamp Comp.

Comoonent 120 160 200. . 120 160 200 120 160 200

Plenum Air. 0060 .0065 .0058 .0064 .0065 .0062 .0080 .0079 .0066

Slabs .6218 .5731 .6055 .6224 .6126 .6138 .5899 .5766 .6279

North Rm. .0189

Wall

.0186 .0170 .0162 .0167 .0171 .0185 .0155 .0143

North .0135

Plenum Wall

.0144 .0125 .0150 .0153 .0141 .0166 .0164 .0147

East Room .0197

Wall

.0226 .0210 .0189 .0197 .0215 .0170 .0206 .0177

East .0139

Plenum Wall

.0150 .0134 .0128 .0136 .0130 .0138 .0148 .0135

South .0435

Room Wall

.0471 .0445 .0403 ‘.0410 .0431 .0382 .0433 .0395

South .0284

Plenum Wall

.0329 .0299 .0292 .0286 .0300 .0317 .0316 .0282

West Room .0158

Wall

.0174 .0191 .0128 .0149 .0152 .0159 .0178 .0146

West .0128

Plenum Wall

.0154 .0141 .0126 .0114 .0122 .0147 .0163 .0130

Lumin- .0792

aires

.0999 .0939 .0770 .0816 .0842 .0669 .0725 .0696

Struct- .1167

ural Steel

.1266 .1138 .1258 .1275 .1206 .1558 .1538 .1295

Suspended .0098 .0106 .0095 .0105 .0107 .0101 .0130 .0129 . 0 1 Ceiling

71
As would be expected, most of the heat storage occurs in the slabs,

typically about 60 percent. The room walls combined account for only

ten percent, while the steel supports in the plenum store twelve

percent, the luminaires eight percent, and the plenum walls seven

percent. The suspended ceiling accounts for only one percent of the

heat storage, and the plenum air one half of one percent.

6.4 Plenum Cooling Load Fraction

Designers frequently want to know the amount of cooling load due to heat

picked up in the room versus the plenum. Heat picked up in the plenum

does not affect room comfort conditions. The less heat which enters the

room, the lower the airflow rates and higher the air temperatures can be

to maintain comfort.

The plenum heat pickup fractions were determined from the ratio of the

temperature rise in the air entering the plenum to the return air

leaving the room, to the temperature rise from the supply to return.

That is

T - T „ Return Room Plenum Fraction = — ——- — — (6.11)

Return Supply

These temperatures were measured with thermocouples, leading to some

scatter in the data. The room air temperature was used as the

temperature of the air leaving the room and entering the plenum. This

is a reasonable assumption, since the air in the test room is well mixed

with the fan on. In addition, for the side slot and lamp compartment

returns, return air enters the plenum at many distributed locations,

making it difficult to measure.

Figure 6.76 shows the plenum heat pick up fraction versus airflow rate

and return air configuration. The values are the average of the plenum

fractions for each of the test conditions, excluding the first several

hours after the lights were turned on, when the scatter in the data is

most prevalent. With the exception of the lamp compartment return, 200

cfm airflow rate condition, the average plenum fraction values ranged

from 0.55 to 0.59. The high plenum fraction at the 200 cfm, lamp

compartment condition is due to the large heat pickup by the return air

being drawn through the lamp compartment at a high flow rate. In

general, the plenum fraction values increased with airflow rate, except

for the ceiling grill return. The plenum fraction with the side slots

increased moderately with airflow rate.

The plenum fraction values from the individual test conditions are

plotted in figure 6.77, 6.78 and 6.79 for the ceiling grill, side slot

and lamp compartment configurations, respectively, all for an airflow

rate of 160 cfm. After the lights are turned on, the plenum fraction

rises quickly, then gradually decreases for about four hours. This may

be due to the gradual heating up of the plenum space. As the plenum

temperatures rise, more heat is lost through the ceiling slab, and

transferred to the room through the suspended ceiling.

72
WATTS

Figure 6 .

1

Measured Cooling Load and Lighting Power for Back-to-Back

Lights On and Lights Off Tests

EM, PARA, 160 CFM, 75 F, LAMP COMP

THIE.HOCRS MEAS COOL UGWT1NG

EM, PARA, 160 CFM, 75 F, U\MP COMP

Figure 6.2 Adjusted Cooling Loads

73
WATTS

WATTS

Figure 6.3 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, CG160 On

EW.PARA, 160 CFM, 75 F, CEIL GRILL

A-O.46464 B-O.1663*

EW.PARA, 160 CFM. 75 F, CEIL GRILL

INJTU4. TRANSIENT

Figure 6.4 Calculated Minus Measured Cooling Load, CG160 On
WATTS

WATTS

Figure 6.5 Measured Cooling Load Versus Double Exponential Regression

Fit, CGI 60 On

EW.PARA, 160 CFM, 75 F, CEIL GRILL

A-0. 46464 B-0. 16634 C-041984 0-1 .21766

EW, PARA, 160 CFM, 75 F, CEIL GRILL

A—0.62233 9-0.20377

. TIME. HOURS

kCAS COOL LIGHTING CALC COOL

Figure 6.6 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, CG160 Off

75
Figure 6.7 Initial Transient, CG160 Off

EW, PARA, 160 CFM, 75 F: CEIL GRILL

INITIAL TRANSIENT

EW, PARA, 160 CFM, 75 F, CEIL GRILL

A-0.62233 B-0.20J77 C-O-38232 0-2.24O6&

time, hours

MEAS COOL —— DOTTING ——- CALC COOL

Figure 6,8 Measured Cooling Load Versus Double Exponential Regression

Fit, CGI 60 Off

76
Figure 6.9 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, CG120 On

EX, PARA, 120 CFM, 75 F, CEIL GRILL

A—0.72791 &-0.24S14

TIME.HOURS

MEAS COOL LIGHTING CALC COOL

EX, PARA, 120 CFM, 75 F, CEIL GRILL

A-0.72791 8—0.2 A6 14 C-O.14638 0-4 250

TIME, HOURS

fcCAS COOL UGHT1NG CALC COOL

Figure 6.10 Measured Cooling Load and Lighting Power Versus Double

Exponential Regression Fit, CG120 On

77
WATTS

Figure 6.11 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, CG200

EV, PARA, 200 CFM, 75 F, CEIL GRILL

A-0.56 150 9-0.19667

TIME. HOURS

UEAS COOL LIGHT)NO — CALC COOL

EV, PARA, 200 CFM, 75 F, CEIL GRILL

Figure 6.12 Initial Transient, CG200

78
WATTS

WATTS

Figure 6

Figure

.13 Measured Cooling Load Versus Double Exponential Regression

Fit, CG200

EV, PARA, 200 CFM, 75 F, CEIL GRILL

A-0.561S0 B-0. 18567 C-OJ9523 0-2.04771

T1UE.H0LKS

ig’ic COOL UGHTING CALC COOL

EU, PARA, 120 CFM, 75 F, SIDE SLOTS

A—0.75126 S-0 24643

HME.HOURS — MEAS COOL UGHT1NO CALC COOL

6.14 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, SS120 On

79
WATTS

WATTS

Figure 6.15 Initial Transient SS120 On

EU, PARA, 120 CFM, 75 F, SIDE SLOTS

INITIAL TRANSIENT

Figure

EU, PARA, 120 CFM, 75 F, SIDE SLOTS

A-0. 75126 8-0.245*3 C-0-236S9 D-4 .03663

6.16 Measured Cooling Load Versus Double Exponential

Fit, SS120 On

Regression

30
Figure 6.17 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, SS120 Off

EU, PARA, 120 CFM, 75 F, SIDE SLOTS

A-0.62008 &-01WSJ

EU, PARA, 120 CFM, 75 F, SIDE SLOTS

INTTUA. TRANSIENT

Figure 6.18 Initial Transient, SS120 Off

SI
WATTS

WATTS

Figure 6.19 Measured Cooling Load Versus Double Exponential Regression

Fit, SS120 Off

EU, PARA. 120 CFM, 75 F, SIDE SLOTS

4-0.62008 B—0.18963 C-0.51963 D

2.20924

900

800

700

800

500

400

300

200

100

0

-100

0 4 8 12 IS 20 24

TIME, HOURS

MEAS COOL ——• CALC COOL

ER, PARA., SIDE SLOTS, 160 CFM, 75 F

4-0.7338 B—0.24967

Figure 6.20 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, SS160 On

82
Figure 6.21 Initial Transient, SS160 On

ER„ PARA., SIDE SLOTS, 1 60 CFM, 75 F

INITIAL TRANSIENT

TIME, HOURS —— CALC-hCAS COOUNO

ER, PARA., SIDE SLOTS, 160 CFM, 75 F

A ^3. 7338 8-0.2*967 0-0.37*66 D-«.S637

TIME, HOURS

MEAS COOL CALC COOL

6. 22 Measured Cooling Load Versus Double Exponential Regression

Fit, SS160 On

33
WATTS

WATTS

Figure 6.23 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, SS160 Off

ER, PARA., SIDE SLOTS, 160 CFM, 75 F

*-0.63026 0-0.1 9671

time, hours

MEAS COOL CALC COOLING

ER, PARA., SIDE SLOTS, 160 CFM, 75 F

INITIAL TRANSIENT, LIGHTS OFF

280

260 240

220 200

180

160

140

120

1008060

4020

0 -20

0 0.2 (3. 4 0 6 0.8 1 1.2 1.4 16

-TTBT HOURS

MEAS—CALC COOUNG

Figure 6.24 Initial Transient, SS160 Off

84
WATTS

WATTS

Figure 6.25 Measured Cooling Load Versus Double Exponential Regression

Fit, SSI 60 Off

ER, PARA., SIDE SLOTS, 160 CFM, 75 F

A—0.63826 0-a 19671 C-0.43316 D-2.45647

TIME. HOURS

MEAS COOUNO CALC COOUNG

ET, PARA, 200 CFM, 75 F, SIDE SLOTS

A—0.6421 39 0-0.23166

Figure 6.26 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, SS200 On

05
WATTS

Figure 6.27 Initial Transient, SS200 On

ET, PARA, 200 CFM, 75 F, SIDE SLOTS

2BO

260

240

220

200

160

160

140

120 10060

60

40

20

0

-20

0 0.2 0.4 0.6 Oa 1 1.2 'HME.HOURS

INITIAL TRANSIENT

ET, PARA, 200 CFM, 75 F, SIDE SLOTS

A-0.642136 8-0.23166 C-O.25632 0-2.6611

Figure 6.28 Measured Cooling Load Versus

Fit, SS200 On

86

Double Exponential Regression
WATTS

Figure 6.29 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, SS200 Off

aoo

700

eoo

aoo

400

300

200

100

0

-100

0 4 a 12 16 20 24 28

TIME.HOURS

MEAS COOL CMJZ COOL

ET, PARA, 200 CFM, 75 F, SIDE SLOTS

A-0. 57479 8-0.21544

%

ET, PARA, 200 CFM, 75 F, SIDE SLOTS

320

300 280

260 240

220

200

180

160

140 120

10080

60

40

20

0

0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8

TIM E.H OURS

Figure 6.30 Initial Transient, SS200 Off

37
Figure 6.31 Measured Cooling Load Versus Double Exponential Regression

Fit, SS200 Off

ET, PARA, 200 CFM, 75 F, SIDE SLOTS

A"0.57*79 8-0.215*4 C-O.3658 D-1 .5832

EJ, PARA, 120 CFM, 75 F, LAMP COMP

A—0. 476D3 8-0.1 1*«3

Figure 6.32 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, LC120 Off

SC
WATTS

Figure 6.33 Initial Transient, LC120 Off

EJ, PARA, 120 CFM, 75 F, LAMP COMP

INITIAL TRANSIENT

EJ, PARA, 120 CFM, 75 F, LAMP COMP

A-0.476M B-0.1146J C-0.52C9« 0-1 .2J763

Figure 6.34 Measured Cooling Load Versus Double Exponential Regression

Fit, LC120 Off

39
WATTS

Figure 6.35 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, LC160 On

EM, PARA, 160 CFM, 75 F, LAMP COMP

EM, PARA, 160 CFM, 75 F, L7\MP COMP

INtTIAL. TRANSiENT

Figure 6.36 Initial Transient, LC160 On

90
WATTS

WATTS

Figure 6.37 Measured Cooling Load Versus Double Exponential Regression

Fit, LC160 On

EM, PARA, 160 CFM, 75 F, LAMP COMP

*-0.62*26 B—O, 18*3* C-0-30737 0-208*61

TIME. HOURS — kCAS COOL LIGHTING CALC COOL

EM, PARA, 160 CFM, 75 F, LAMP COMP

Figure 6.38 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, LC160 Off

91
Figure 6.39 Initial Transient, LC160 Off

EM, PARA, 160 CFM, 75 F, U\MP COMP

INITIAL TRANSIENT

EM, PARA, 160 CFM, 75 F, LAMP COMP

A-O.S8570 9-0.18968 C-0.391S0 0-1 86102

TIME.HOURS

MEAS COOL — CALC COOL

Figure 6.40 Measured Cooling Load Versus Double Exponential Regression

Fit, LC160 Off

92
WATTS

Figure 6.41 Measured Cooling Load and Lighting Power Versus Single

Exponential Regression Fit, LC200 Off

EN, PARA, 200 CFM, 75 F, L.AMP COMP

Q—0.601 3*EXP(—0.2229*T)

— TIME. HOURS

MEAS COOUNG CALC COOLING

EN, PARA, 200 CFM, 75 F, LAMP COMP

INITIAL. TRANSIENT

ICAS-CALC COOLWG

Figure 6.42 Initial Transient, LC200 Off

93
Figure 6.43 Measured Cooling Load Versus Double Exponential Regression

Fit, LC200 Off

EN, PARA. 200 CFM, 75 F. LAMP COMP

A—0.601 3 0-0.2229 C-0 56 12 0-241033

— CMX ——• hEAS

ES, ET LIGHTS ON VS OFF

200 CFM, 75 F, SICE SLOTS

TJME.HOURS Q ETON + £T OFF O ES ON A ES OFF — AVG

Figure 6.44 Comparison of Cooling Load Profiles From Different Tests

with Same Conditions

94
REGRESSION

COEFFICIENTS

Figure 6.45 Regression Coefficients Versus Airflow Rate

REGRESSION COEFFICIENTS VS AIRFLOW

4.5

4

3.5

3

2.5

2

1.5

1

0.5

0 100 120 140 160 180 200 220

NRFLOW, CFU 0 A T S O C & D

ALL RETURN NR PATHS

REGRESSION COEFFICIENTS VS AIRFLOW

N_L RETURN NR PATHS

0.8 -

0.7 - a

0.6 - a 8

a

5

kiu 0.5 -

k.E 0

0u 0.4 - •

z *

2

\A

V)

ki 0.3 -

R o

o

5 X

ki 4

c 4 4 4

0.2 -

+

4

« X

0.1 - 4 X

5

X X

X X

i i '

I I I I i —"I ' I" i I 100 120 140 160 180 200 220

NRELOW, CFM

A -* 0 R C

Figure 6.46 Regression Coefficients A, B and C Versus Airflow

$5
REGRESSION

COEFFICIENTS

REGRESSION

COEFFICIENTS

Figure 6.47 Regression Coefficient A Versus Airflow

REGRESSION COEFF A VS AIRFLOW

ALL RETURN AIR PATHS

0.74

0.72

0.7

0.68

0.66

0.64

0.62 06

0.58

0.56

0.54

0.52

0.5

0.48

0.46 100 120 140 160 180 200 220

o CL G«_

AIRFLOW. CFM

T S SLTS o LHP COMP

REGRESSION COEFF B VS AIRFLOW

ALL RETURN AIR PATHS

0.25

0.24

0.25

0.22

0.2

1

0.2

0.19

0 18

0.17

0 16

0.15

0.14

0.13

0.12

0.1

1

100 120 . 140 160 180 200 220

G CL GRL

AIRFLOW, CFM

S SLTS o LMP COMP

Figure 6.48 Regression Coefficient B Versus Airflow

96
REGRESSION

COEFFICIENTS

REGRESSION

COEFFICIENTS

Figure 6.49 Regression Coefficient C Versus Airflow

REGRESSION COEFF C VS AIRFLOW

0.55

0.6

0.46

0.4

0-35

0.3

0 26

0.2

0.16

0.1 100 120 140 160 180 200 220

MRFLOW, CFM

CL GRL +• S SLTS O LMP COUP

ALL RETURN MR PATHS

REGRESSION COEFF D VS AIRFLOW

4.4

4 2

4

3.8

3.6 3 4

3.23

2.8

2.6

2.4

2.2

2

1.8

1.6

1.4

1.2 100 120 140 160 180 200 220

AIRFLOW, CFM

CL G«_ + S SLTS © LMP COUP

Figure 6.50 Regression Coefficient D Versus Airflow

ALL RETURN MR PATHS

97
COOLING

LOLD/LIGHTING

POWCft

Figure 6.51 Cooling Load Profiles for Ceiling Grill Return

COOLING LOAD PROFILE VS AIRFLOW

CEILING GRILL RETURN

TTME.HOURS Q 120 CFU + 160 CFW © 200 CFVI

COOLING LOAD PROFILE VS AIRFLOW

SIDE SLOT RETURN

1 20 cm

TTME.H0URS

160 cm 2oo cm

Figure 6.52 Cooling Load Profiles for Ceiling Side Slot Return

98
COOLING

LOAD/

LIGHTING

POWER

COOLING

LOAD/LIGHTING

POWER

Figure 6.53 Cooling Load Profiles for Ceiling Lamp Compartment Return

COOLING LOAD PROFILE VS AIRFLOW

LAMP COMPARTMENT RETURN

TIME, HOURS

120 CFM + 160 CFM 0 200 CFM

COOLING LOAD PROFILE VS RETURN AIR PATH

C GRILLS SLOTS OR LMP CMP. 120 CFM

TIME, HOURS

Q CEL GRL * -* S SLOTS O LMP COMP

Figure 6.54 Cooling Load Profiles for Ceiling 120 cfm Airflow
COOLING

LO*0/LIGHTING

POWER

COOLING

LCMO/UGHTING

POWER

Figure 6.55 Cooling Load Profiles for Ceiling 160 cfm Airflow

COOLING LOAD PROFILE VS RETURN AIR PATH

C GfWLLS SLOTS OR LMP CMP. 160 CFM

Q CEC 6RL

TIME, HOURS

+ S SLOTS P LMP COMP

COOLING LOAD PROFILE VS RETURN AIR PATH

C GRILLS SLOTS OR LMP CMP, 200 CFM

T1ME.H0URS O CEIL GRL •+ S SLOTS 9 LMP COMP

Figure 6.56 Cooling Load Profiles for Ceiling 200 cfm Airflow

100
WEIGHTING

rACTOKS

COOLING/LIGHTING

Figure 6.57 Cooling Load One Hour After Lights are Energized

COOLING LOAD AFTER 1 HOUR

LAMP COMP. SCE SLOTS OR CEIL GRIL

AIRFLOW RATE. CFM LMP CMP S SLOTS O CEL GRL

WEIGHTING FACTOR AO VS AIRFLOW

ALL RETURN AIR PATHS

Figure 6.58 Weighting Factor aQ , Versus Airflow

101
WEIGHTING

FACTORS

WEIGHTING

FACTORS

Figure 6.59 Weighting Factor a^ Versus Airflow

0.65

0.6

0.56

0.5

0.45

0.4

QJ5

0.3

0.25

0.2 100 120 140 160 180 200 220

a CL GRL

AIRFLOW, CFH

* S SLTS o LMP COkP

WEIGHTING FACTOR A1 VS AIRFLOW

ALL RETURN MR PATHS

WEIGHTING FACTOR A2 VS AIRFLOW

ALL RETURN AIR PATHS

Figure 6.60 Weighting Factor a 2 Versus Airflow

102
WEIGHTING

FACTORS

Figure 6.61 Weighting Factor Versus Airflow

WEIGHTING FACTOR B1 VS AIRFLOW

ALL RETURN AIR PATHS

AIRFLOW, CPU Q CL GRL S SLTS O LMPCOMP*

M

XO

S

2

(3Z

r0

1

WEIGHTING FACTOR B2 VS AIRFLOW

ALL RETURN AIR PATHS

AIRFLOW, CFU O CL OR. + S SLTS O LMP CO»rf>

Figure 6.62 Weighting Factor \>2 Versus Airflow

103
COOLING/

LIGHTING

COO

LING/

LIGHTING

Figure 6.63 Hourly Cooling Load Fractions for Four Days

COOLING LOAD FRACTION, 10 HOURS ON

CERJNG GRILL 120 CFM

1

0.9

as

0.7

0.6

as

0.4

0.3

0.2

I

0.1

0

0 4 8 12 16 20 24

TIME, HOURS

CEILING GRILL 120 CFM

COOLING LOAD FRACTION, 10 HOURS ON

CEIUNG GRIX

1

-

0.9

0.8

0.7

0.6

O.S

0.4

0.3

0.2

I

0.1

0

0 4 8 12 16 20 24

TIME. HOURS

120 CFM 160 CFM O 200 CFM

Figure 6.64 Hourly Cooling Load Fractions Versus Airflow Rate for

Ceiling Grill

CEIUNG GRIX

104
COOLING/LIGHTING

COOLING/

LIGHTING

Figure 6.65 Hourly Cooling Load Fractions Versus Airflow Rate for Side

Slots

COOLING LOAD FRACTION, 10 HOURS ON

TIME, HOURS

120 CFM + ISO CFM 9 200 CFVI

COOLING LOAD FRACTION, 10 HOURS ON

LAMP COMPARTMENT

1

0.9

0.0

0.7

0.6

0.5

0.4

0.3

0.2 I

0.1

0

o * a 12 16 20 24

TIME, HOURS

120 CFM + ISO OEM O 200 CFM

Figure 6.66 Hourly Cooling Load Fractions Versus Airflow Rate for Lamp

Compartment

LAMP COMPARTMENT

105
COOLING

LOAO.

WATTS

COOLING/

L

IGMT

ING

Figure 6.67 Hourly Cooling Load

Average Conditions

Fractions Versus Airflow Rate for

COOLING LOAD FRACTION, 10 HOURS ON

AVERAGE VCGHTUG FACTORS

1

0.9

0.6

0.7

0.6

0.5

0.4

0.3

0.2

I

0.1

0 0 4 8 12 16 20 24 me. hours

COOLING LOAD PROFILE, 10 HOURS ON

CESJNG GRILL RETURN

900

700

600

500

400

300

200

100 1

0

0 4 8 12 16 20 24

RUE. HOURS

O 12D CFN 160 CFU O 200 CFU

Figure 6.68 Cooling Load Profiles Versus Airflow for Ceiling Grill

AVERAGE VCCHnNG FACTORS

106
COOLING

LOAD.

WATTS

COOLING

LOAD.

WATTS

Figure 6.69 Cooling Load Profiles Versus Airflow for Side Slots

COOLING LOAD PROFILE, 10 HOURS ON

TIME, HOURS

120 CFH + 160 CFH ©200 CFU

COOLING LOAD PROFILE, 10 HOURS ON

LAMP COMPARTMENT RETURN

TIME. HOURS

120 CRM + ISO CFM © 200 CFM

Figure 6.70 Cooling Load Profiles Versus Airflow for Lamp Compartment

107
COOLING

LOAD.

WATTS

COOLING

LOAD.

WATTS

Figure 6.71 Cooling Load Fractions with Original and Adjusted Weighting

Factors, CG200

COOLING LOAD PROFILE, 10 HOURS ON

ADJUSTED WEIGHTING FACTORS. CC200

TIKE. HOURS

ORKJ T NEW A2 © NEW A1

Figure

COOLING LOAD PROFILE, 10 HOURS ON

ADJUSTED WEIGHTING FACTORS, SSI 20

j - - _______ ___ - - T '

0.9 -

0.8 -

0.7 -

0.6 -

0.5 -

o.a a

+

0.3 5

9

0.2 2

SI

0.1

0 - - _____

_Q.i -|—

—j— —

f 1—-—i— r —r

1 1 1 1 1 —

O 4 . 8 12 16 2D 24

TIME, HOURS

ORIG +• NEW A1

6.72 Cooling Load Fractions with Original and Adjusted Weighting

Factors, SS160

ADJUSTED WEIGHTING FACTORS, SSI 2D

" CT

+ 9

?

t “T 1 1 1 1 1— 1 1 1 r

108
watt-hr/Watt

cooling

load,

watts

Figure 6.73 Cooling Load Fractions with Original and Adjusted Weighting

Factors, SS160

COOLING LOAD PROFILE, 10 HOURS ON

ADJUSTED WEIGHTING FACTORS. SS200

1.1 -i — — — —

, .

• O

0.9

0.8

0.7

0.6

0.8

0.4

0.3

0.2

| 0.1

0 a

_o.i _j— 1 | 1 1 1 1 1 1 1 1 1 0 4 8 12 16 20 24

TIME. HOURS

O 0R1G + NEW A1

TOTAL HEAT STORED PER LIGHTING WATT

LAMP COMP, SCE SLOTS OR CEIL GRLL

4.6

4.5

4.4 4 3

4.2

4.14

3.9

3.8

3.7

3.6

3.5

3.4

3.3

3.2

3.13

2.9 2.8 100 120 140 160 180 200 220

AIRFLOW RATE. CFM LMP CMP S SLOTS © CEL ORL

Figure 6.74 Total Heat Storage Versus Airflow Rate and Return

Configuration

ADJUSTED WEIGHTING FACTORS. SS200

a

t

: 1 1 1 1 1 1 1 r

109
PLENUM

FRACTION

WATTS/SQM

Figure 6.75 Heat Flow Sensor Readings for Floor and Walls During Backto-Back Lights on and Lights off Tests

8

765

43

2

10 -f-2

-3

—4

-5

-fi -7

-8

-9

0 2D 40 60 80

TIME. HOURS

PLENUM HEAT PICKUP FRACTION

LAMP COMP, SCE SLOTS OR CEIL Gfill 0.67

0.66

0.65

0.64

0.63

0 62

0.61

0.6

0.59

a.sa

0.57

0.56 0&6

100 120 . 140 160 180 20Q 220

AIRFLOW RATE, CFM Q LMP CMP + S SLOTS « CEL GRL

Figure 6.76 Plenum Heat Pickup Fraction Versus Airflow Rate and Return

Air Configuration

ER, PARA, ISO CFM 75 F

SURFACE HEAT FLOWS

110
PLENUM

FRACTION

Figure 6.77 Plenum Fraction Versus Time for 160 cfm, Ceiling Grill

Return

EW, PARA, 160 CFM, 75 F, CEIL QRILL

PLENUM FRACTION

0.7 -r— — — —— -

0.66 -

0.46 -

0 ** I

1——i1 1 1 1 i —i 1 1——r——r—

t- 6 IQ 14 18 22 26 30 34

THE, HOURS

ER, PARA, 160 CFM, 75 F^SlDt 5 LOTS

PLENUM FRACTION (AOJ) 0.72

0.7 0.66

0.66

0.64

0.62

0.6 056

0.56

0.54

0.52

0.5 0.46 0.46

0.44 0.42

0.4 *2 16 20 24 28 32 36 40 44 46

THE, HOURS

+ AVO Figure 6.78 Plenum Fraction Versus Time for 160 cfm, Side Slot Return

111
PLENUM

Figure 6.79 Plenum Fraction Versus Time for 160 cfm, Lamp Compartment

Return

EM. PARA, 160 CFM, 75. F, LAMP COMP

PLENUM FRACTION

0.9

0.6

0.7

0.6

0.5

0.4 14 18 22 26 30 34 38 42 T1ME.H0URS

112
7. Lighting Energy Distribution Fractions

The distribution of the energy supplied to the lighting system by the

various energy transfer mechanisms was determined from computer

simulations of the test facility. The simulation results were compared

to the corresponding measured data, and the computer model adjusted

until cooling loads and temperatures matched. The lighting energy

distribution fractions were determined from the corresponding values

calculated by the computer model.

Of the nine lighting energy distribution fractions listed earlier, one

is a constant, the ballast fraction, and two are zero for the luminaires

tested. These two are visible light up and convection to return air.

All of the visible light is directed downward, so the upward visible

fraction is zero. The return air is not ducted at the luminaire, so the

convection to return air fraction is zero.

The visible light down fraction is determined from the ratio of lighting

visible watts to lighting power, which are functions of lamp

temperature. The infrared upward and downward fractions were determined

by summing the net radiation exchange between the luminaire nodes and

the room nodes, and plenum nodes respectively. The convection downward

and upward fractions were determined by summing the convective exchange

between the luminaire nodes and the room nodes, and plenum nodes,

respectively. The heat stored fraction is determined from the

difference between the lighting power input and the sum of the other

energy distribution, amounts

.

Simulations were run for two return air configurations, either lamp

compartment or ceiling return. The ceiling return corresponds to either

the ceiling grill or side slot return, which the measurement results

showed performed similarly. For the ceiling return, convective heat

transfer from the lower luminaire housing surface and the lamps was

assigned to the room air. For the lamp compartment return all

convection from the luminaire was directed into the plenum air. Two

airflow rates were simulated for each return configuration, 120 and 200

cfm. Also, the effects of a different room air temperature and double

the amount of floor slab thermal mass were simulated.

7.1 Calibration and Validation of Computer Model

The calibration of the computer model involved adjusting the convective

heat transfer coefficients so that the calculated surface and air

temperatures matched the measured values. The initial assumption was

free convection for each surface node. The calculated surface

temperatures were then compared to the corresponding average measured

values. If the calculated surface temperature for a node was too high,

the convection coefficient was increased until the calculated and

measured temperatures matched. The computer model was individually

calibrated versus measured results for each simulation configuration.

The calculated cooling load profile was compared to the measured

profile. Figure 7.1 shows a comparison of the measured and calculated

cooling load profiles for a lights-on step response test, for ceiling

return at 120 cfm airflow rate. In general, simulated and measured

113
values are in good agreement although, the simulated cooling load

exceeds the measured during the first four hours after lights on.

Figure 7.2 presents a similar comparison of measured to simulated

cooling load for the ceiling return, 200 cfm condition. Very good

agreement is seen between the measured and simulated values.

For the lamp compartment return configuration simulations, the measured

and simulated cooling loads are compared in figures 7.3 and 7.4 for 120

cfm and 200 cfm airflow rates, respectively. The agreement between

measured and simulated cooling loads is poorest for the 120 cfm

condition, but best for the 200 cfm condition.

7.2 Computer Simulation Results

The lighting energy distribution fractions are shown as a function of

time in figure 7.5, for the ceiling return, 120 cfm condition. The

visible down fraction is nearly constant at 18 percent. The greatest

heat transfer from the luminaire is by convection to the room air, about

42 percent, The downward convection fraction starts out at about 35

percent and increases over the first four hours. The downward infrared

heat transfer constitutes 26 percent of the lighting energy input. This

quantity, and the upward infrared heat transfer of about 9 percent,

start out at their maximum values and gradually decreases until they

reach steady state values. The smallest lighting energy distribution

fraction is the upward convection, which barely exceeds 5 percent.

The distribution fractions are quite a bit different for the lamp

compartment return configuration, as shown in figure 7.6. Heat transfer

to the plenum air by convection becomes the dominant mode of luminaire

heat dissipation, topping 60 percent. Downward infrared heat exchange

drops to 16 percent, just below downward visible light at 18 percent.

Upward infrared heat transfer is cut in half to four percent.

Similar lighting energy fractions were determined for airflow rates of

200 cfm, room air temperature of 80°F, and double the thermal mass in

the slabs. Figure 7.7 shows the lighting energy distribution fractions

for the ceiling return, 120 cfm airflow rate with a room air temperature

of 80°F. These results are very similar to the same configurations at

75°F’ except downward infrared radiation is greater by two percent.

This indicates that room air temperature set - point may not influence

the lighting energy distribution fractions, for room temperatures

typically encountered.

Figure 7.8 shows the simulated cooling load profile for the test room if

the mass of the floor and ceiling slabs were doubled. Also plotted, for

comparison, are the measured cooling load profiles for the ceiling grill

return and airflow rates of 120 and 200 cfm. The additional thermal

mass provides for much greater heat storage, and the potential for much

lower peak cooling loads. After the lights have been on for 10 hours,

cooling loads would only be 75 percent of lighting power for the heavy

mass room, versus 90 percent of lighting power for the baseline room, a

peak cooling load reduction of 17 percent. Figure 7.9 shows the

lighting energy distribution fractions for the heavy thermal mass

condition, with ceiling return and 120 cfm airflow. These values are

very similar to those for the normal mass room.

114
The values of the lighting energy distribution fractions at equilibrium

conditions are summarized in table 7.1 for the two airflow rates and

return air configurations, the double mass configuration and the 80°F

condition.

Table 7.1 Lighting Energy Distribution Fractions

Energy Ceiling Return Lamp Compartment Heavy

Distribution • Return Thermal

Mechanism 120cfm 200cfm 120cfm 200cfm Mass oo o

o

Visible down 0.1783 0.1782 0.1778 0.1777 0.1783 0.1786

Visible up 0 0 0 0 0 0

Infrared down 0.2568 0.2455 0.1656 0.1511 0.2576 0.2608

Infrared up 0.0944 0.0903 0.0427 0.0454 0.0944 0.0962

Convection down 0.4160 0.4269 0 0 0.4133 0.4105

Convection up 0.0545 0.0591 0.6139 0.6260 0.0546 0.0539

Convection return 0 0 0 0 0 0

The visible down fraction remains nearly constant. As would be

expected, the upward convective fraction increases with airflow rate,

and is greater for the lamp compartment return. The increases in upward

convection are at the expense of downward convection, and upward and

downward infrared radiation exchange. Upward convection is less than

six percent with the ceiling return, but exceeds 60 percent with for the

lamp compartment return. The downward infrared fraction is lower by 36

percent and the upward infrared fraction lower by 50 percent, for the

lamp compartment return than the ceiling return.

115
COOLING

LOAD

FRACTION

COOLING

LOAD

FRACTION

Figure 7 .

1

Comparison of Measured and Simulated Cooling

Loads, Ceiling Return, 120 cfm

MEASURED VS SIMULATED COOLING LOAD

CSUNG RETURN, 120 CFM, 75 F

O SI MOL COOL

TIM&HOURS + MCAS COOL

MEASURED VS SIMULATED COOLING LOAD

Figure 7.2 Comparison of Measured and Simulated Cooling

Loads, Ceiling Return, 200 cfm

116
COOLING

LOAD

FRACTION

Figure 7.3 Comparison of Measured and Simulated Cooling

Loads, Lamp Compartment Return, 120 cfm

MEASURED VS SIMULATED COOLING LOAD

LAMP COMPARTMENT. 120 CFM. 75 F

SIMUL COOL

TIME, HOURS

MEAS COOL

MEASURED VS SIMUATED COOLING LOAD

LAMP COMPARTMENT, 200 CFM, 75 F

TIME, HOURS

O SIMUL COOL + MEAS COOL

Figure 7.4 Comparison of Measured and Simulated Cooling

Loads , Lamp Compartment Return 200 cfm

117
LIGHTING

ENEHGT

FRACTION

LIGHTING

ENERGY

FRACTIONS

Figure 7.5 Lighting Energy Distribution Fractions Versus

Time, Ceiling Return, 120 cfm

LIGHTING ENERGY FRACTIONS

CEUNG RETURN, 1 2D CFM

TIME, HOURS

Vt> T IRO O HU & COO X COU

LIGHTING ENERGY FRACTIONS

LAMP COMPARTWENT, 1 20 CFM, 75 F

Figure 7.6 Lighting Energy Distribution Fractions Versus

Time, Lamp Compartment Return, 120 cfm

118
LIGHTING

ENCnarr

Figure 7.7 Lighting Energy Distribution Fractions Versus

Time, Ceiling Return, 120 cfm, 80°F

LIGHTING ENERGY FRACTIONS, 80 F

CEUNC RETURN. 1 20 CFM

a vo + iro o «u a coo x cou

COOLING LOAD PROFILE, HEAVY MASS

DOUBLE SLAB MASS

Figure 7.8 Cooling Load Profile for Heavy Thermal Mass

Room, Ceiling Return, 120 cfm

119
LIGHTING

ENERGY

FRACTIONS

LIGHTING ENERGY FRACTIONS

HEAVY MASS. CEIUNG RET. 120 CFVI

TIUEHOLWS

o vo iro o wu a coo x cou

Figure 7.9 Lighting Energy Distribution Fractions Versus

Time, Heavy Thermal Hass Room, Ceiling Return,

120 cfm

120
8 . Discussion of Results

Examination of the measurement and simulation results completed thus far

reveals a series of preliminary general results related to the

interaction of the lighting and HVAC system, and the room. The

similarities and differences between the results for the different test

conditions help to illuminate the factors which contribute to the energy

and thermal performance of lighting systems in actual room environments.

This information, in turn, will have potential impact on design

considerations for lighting and HVAC systems.

One point is that although massive room components such as the floor

slab may respond slowly to the effects of the lighting system, the

luminaires themselves reach equilibrium relatively quickly, in two to

three hours. Thus, the airflow conditions in the room should be

selected so that the luminaires are at their most efficient operating

point. For the type and size of the luminaire tested, lamp compartment

return was required to achieve this. For 2-tube luminaires instead of

4- tube, the side slot or ceiling grill return may have been sufficient

at high enough airflow rates. Two-tube luminaires are scheduled for

future testing.

Lamp compartment return cools the lamps significantly more than the side

slot or ceiling grill return. The performance with the side slots and

ceiling grill were very similar. It may be adequate to lump these two

return configurations into a single category for design considerations.

Most of the heat storage, about 60 percent was in the floor slabs. The

portions of the sidewalls above the suspended ceiling stored about the

same amount of heat as the room walls, their smaller volumes being

compensated by higher equilibrium temperatures with the lights on. The

room walls do not store much heat, due to relatively low thermal mass

and small elevations in temperature above room air temperature.

Convection at the room walls must be sufficiently high to keep the walls

cool. If wall convective heat transfer could be reduced at the walls,

more heat storage could occur. This could be accomplished by keeping

the supply and return air registers away from the walls, where possible.

Lowering airflow rates would also allow more heat storage.

From the point-of-view of controlling peak cooling load, the optimum

cooling load profile due to cyclic operation of the lights would be a horizontal line. The height of the line would equal the product of the

lighting power and the fraction of the day the lights are on. This

minimum peak cooling load to remove the heat from the lights is given

by:

Minimum

Peak = (Lighting Power) (Fraction of Day Lights are on) (8.1)

Cooling

Load

For these tests, lighting power averaged 750 watts. The minimum peak

cooling load for 10 hour cyclic operation is 42 percent of the lighting

power, or 313 watts. This is significantly less than the 90 percent

values observed for the test conditions.

121
In order to make the cooling load profiles for cyclic operation of the

lights determined from these tests into a straight line with zero slope,

the value for the regression coefficient B would have to be smaller, and

the A values larger. Increasing A corresponds to decreasing the initial

cooling load by storing more heat. This must be accomplished while

still cooling the lamps to an efficient operating point. Reducing B

corresponds to increasing long term heat storage, by adding more thermal

mass. Adding furnishings would increase B.

Another way to look at this point would be to to cool the room at high

airflow rates in the morning, then reduce airflow rates in the afternoon

to force more of the lighting energy into storage. In this manner, the

heat storage capacity is restored in the morning for use later in the

day. A side benefit of reducing airflow rates in the afternoon is that

lighting power could go down due to higher lamp temperatures. This is a

double-edged sword though, because visible light output also goes down.

Of the heat storage in the room floor, more of the heat gain comes from

the plenum below then from the room. There is a net flow of heat from

the plenum into the slab above when the lights are turned on, then into

the room when the lights are turned off. An equal amount of heat is

lost through the ceiling slab above. Insulating the room from the

plenum below could reduce the heat gain to the room. This could be

accomplished by adding carpeting and padding. This would allow the

floor and plenum to heat up more, thus storing more heat. Storing heat

in the plenum is desirable since it is an unoccupied space with no

comfort requirements. In contrast, as more heat is stored in the room,

the impacts on thermal comfort should not be overlooked. As radiant

temperatures increase, lower air temperatures or higher airflow rates

are required to maintain equivalent comfort conditions [As77 ]

.

Topics for consideration for future research would include looking at

the effects of carpet, and the effects of different supply air

diffusers. Also, the effects of variable air volume operation on

cooling load profiles would be of interest. The effects of different

luminaire types and different lighting power densities on the steady

state and transient performance of the lighting and cooling systems also

are of interest. Other potential items for testing consideration

include other lamp types and dimming controls.

122
9. Conclusions

This interim report summarizes preliminary results regarding the effect

of interactions between building lighting, heating and cooling systems

on the energy performance of the lighting systems, and cooling loads due

to lighting. The evaluation consisted of full-scale measurements and

related computer simulations.

A unique test facility was designed, constructed and operated to emulate

an office space with recessed fluorescent lighting system, surrounded by

similar spaces on all six sides. The test facility was specially

instrumented to record energy usage, cooling load, heat flows,

temperatures and light levels. Measurements were made every twelve

seconds and averaged over two minute intervals.

Steady state tests of lighting system performance demonstrated the

temperature dependence of the light output and power consumption of the

lighting system, which consisted of four luminaires with four lamps

each. Temperature dependent correlations were derived for light output

and lighting power, based on the measurements. Changes in room air

temperature of 5°F produced 4 to 5 °F changes in room surface

temperatures, and about 3°F change in minimum lamp temperature.

Lighting power and light output varied about five percent for the range

of conditions tested. The return plenum configuration provides some

cooling for the luminaire even at the lowest airflow rate.

Airflow rate was observed to have a modest but predictable effect on

lighting thermal performance, increasing convective transfer from the

luminaire. Return air configuration had a much greater effect,

particularly lamp compartment extract. Minimum lamp temperatures were

an average of seven degrees less for the lamp compartment extract than

either the ceiling grill or side slot returns, meaning greater light

output and higher efficiency. The results for the ceiling grill and

side slots were generally similar.

Results from transient measurements were generally similar for all test

conditions, except for the lamp compartment return. More heat storage

occurred under that configuration, especially at the lowest airflow

rate. ASHRAE-type weighting factors were computed from measurements for

all the test conditions.

A new double exponential curve fit was developed to improve the fit

between the measured cooling load profile and the calculated profile.

The weighting factors corresponding to the new curve fit were derived.

Regression coefficients and extended weighting factors were determined

from measurement data for each test condition. A set of regression

coefficients and weighting factors was determined for the average of all

the test conditions. Calculations using the weighting factors showed

that for 10 hours-on daily cyclic lighting system usage, the steady

periodic peak cooling load would be about 90 percent of lighting power

for the conditions tested. The weighting factors can be used to

determine cooling load due to any arbitrary lighting schedule. For the

conditions tested, approximately 24 hours was required for the room to

reach equilibrium. Doubling the floor mass reduced peak cooling load by

17 percent.

123
Heat storage occurred mainly in the slabs, accounting for approximately

60 percent of total heat storage. The room walls and plenum walls

stored equivalent amounts, as the plenum walls heated up about four

times as much. More heat enters the floor slab from the lower plenum

than the room, and more floor slab heat is discharged into the room than

returns to the plenum below, leaving a net heat gain to the room. This

is matched by a heat loss through the ceiling.

On average, about 57 percent of the total room cooling load was picked

up in the plenum, leaving 43 percent due to room loads. The individual

lighting energy distribution fractions were determined from computer

simulations and tabulated. These values can be used as input values for

building energy analysis computer programs.

The impact of the measurement results on design considerations was

discussed. The lamp compartment return was seen to be very effective at

keeping the lamps at a cool and efficient operating condition.

Operating the fans at high speed in the morning and low speed in the

afternoon was suggested as a way to reduce peak cooling loads.

Reduction of convection at the room walls would allow for more heat

storage in those elements. A carpeted floor would force more heat

storage in the plenum. The plenum is the most desirable heat storage

location, because it has minimal comfort effects. More mass in the

floor or walls would reduce peak cooling loads, as would the addition of

furnishings

.

In closing, preliminary results from this research showed that lighting

systems can be constrained to operate at their most efficient level,

through manipulation of the thermal environment via the air temperature,

airflow rate and return air configuration. A unique test facility was

designed, built and operated, with a movel design to maintain thermal

boundary conditions. The experimental plan, and test facility control

system are original. A unique set of steady-state and transient

measurements was collected and analyzed. A new regression correlation

was developed to improve the fit between measured and calculated cooling

load, and new weighting factors derived. . From analysis of the

measurements, the energy distribution, and storage fractions were

determined, both unique efforts. A detailed computer model was

developed, calibrated and validated. Based on the measurements and

modeling, design recommendations and considerations were developed.

124
Appendix

program lites25

C 25 NODE TRANSIENT FINITE DIFFERENCE HEAT BALANCE MODEL FOR

C -LIGHTING- AND -ROOM -THERMAL -ANALYSIS

C

real e-f (25) ,em(25) ,t (25) ,b (25) ,h (25) ,-f (25,25) ,qrn (25) ,qc (25) , * hc(2S) ,a(25) ,qs(25) ,c(25) ,s(25) ,qv(25) .alpha (25) ,lp,ll

,

* MAXC,SUM(25) ,0B(25) , LO , LPM , LLM , CK (25) , IRD, IRU,LLL,LPL

double precision eb (25), sigma

character*? -fn DO 997 1-1,25

CK ( I ) -0.

DO 997 J-1,25

F ( I , J) —O.

997 CONTINUE

C RADIATION SHAPE FACTORS

F (8,4)—. 0403

F (8,6)—. 0403

F(8, 15) -.263691

F(8, 16)-. 150497

F(8, 17)-. 177356

F(8, 18)-. 150497

F(8, 19)-. 177356

F(ll ,23)-. 081719

F(ll ,20)-. 069157

F(ll ,21)-. 081719

F(ll ,22)-. 069157

F(ll ,25)-. 165078

F ( 1 1 , 14)-. 533169

F(14, 22)-. 077968

F(14, 21)-. 092720

F(14, 20)-. 077968

F(14,23)». 092720

F( 14, 11 >-.658623

F(1S, 8)-. 325736

F (15,16)-. 154689

F ( 15, 17) -. 182443

F(15, 18)-. 154689

F(15, 19)-. 182443

F(16, 4)-. 022113

F(16, 6)-. 022113

F(16, 8)-. 263370

F(16, 15)». 219143

F(16, 17)-. 177692

F(16, 18)-. 117876

F (16,19)—. 177692

F(17, 4)-. 022248

F(17,6)». 022248

F(17, 8)-. 266034

F (17,15)-. 221538

F ( 17, 16) -. 152308

F(17, 18)-. 152308

F(17, 19)-. 163316

F(18, 4)-. 022113

F(18,6)». 022113

F ( 1 8 , 8 ) — . 26337

F(18, 15)-. 219143

F ( 18 , 16) — 117876

F ( 18 , 1 7) — . 177692

F(18, 19)-. 177692

F(19, 4)-. 022248

F(19, 6)-. 022248

F(19, 8)-. 266034

F (19, 15)-. 221538

F(19, 16>«. 152308

F(19, 17)-. 163316

F ( 1 9 , 18)-. 152308

F (20,25) -. 03382

F(20, 11)-. 387280

F(20. 14)-. 353455

125
F (20,21 > = . 092776

F (20,22) *.039893

F (20,23) =.092776

F (21 ,25) =.031964

F (21 , 1 1 >=.392253

F (21 , 14) =. 360284

F (21 ,20) =.079522

F (21 ,22) =.079522

F (21 ,23) =.056455

F (22,25)=. 03382

F( 22, 11)=. 387280

F (22 , 14) =.353455

F (22,20) =.039893

F (22,21 >=. 092776

F (22,23) =. 092776

F (23,25) =. 031964

F( 23, 11)=. 392253

F (23, 14)=. 360284

F (23,20) =.079522

F (23, 21 )=. 056455

F (23,22) =.079522

• ' F (23,25) =.031964

F < 4 , 6 ) = . 5

F (4,8) =. 21 15825

F ( 4 , 16) =.066339

F (4, 17) =. 077868

F ( 4 , 18) =.066339

F (4, 19) =. 077868

F (6,4) =.

5

F (6,8)=. 21 15825

F (6, 16) =. 066339

F (6, 17) =.077868

F ( 6 , 18)=. 066339

F ( 6 , 19)=. 077868

F (25 , 1 1 )=. 866664

F (25,20) =.031707

F (25,21 ) =.034961

F (25,22) =. 031707

F (25,23) =. 034961

C SURFACE AREAS

a ( 1

) =0.

a < 2 ) =0

.

a (3) =0.

a (4) =32.

A (5) =1 . 33

A (6) =32.

A (7) =32.

A (8) =168.

A (9) =168.

A (10) =168.

A ( 1 1 ) =168.

A ( 12) =168.

A ( 13) =168.

A ( 14) =136.

A ( 15) =136.

A ( 16) =96.

A ( 17) =1 12.

A ( 18) =96.

A ( 19) =1 12.

A (20) =30.

A (21 ) =35.

A (22) =30.

A (23) =35.

A (24 ) =30.

A (25) =32.

C STARTING. TEMRERSTURES

126
nn

DO 999 1-1,23

T ( I

) -73.

999 CONTINUE

C THERMAL CAPACITANCE

c < 1 ) -23. 97

c (2) -0. 11 c <3) -. 135

c (4 ) -3.

0

C (5) -6.

0

C (6) — 17.

0

C (7) -1 . 44

C (0) -359.

4

C (9) -339.

4

C ( 10) —359.

4

C ( 1 1

) —359.

4

C ( 12) -339.

C ( 13) -359.

C ( 14 ) — 1 3. 23

C ( 15) — 13. 23

C ( 16) -52.

C(17)— 121.4

C ( 18) -52.

C < 19) -60.

7

C (20) -16.3

C (21 ) -37.

7

C (22) -16.3

C (23) —14.6

C (24) -150.

4

C (25) -0. 49

CONDUCTANCES

CK (5) -12.

CK (6) -12.

CK (0) —070. 91

CK (9) -070. 91

CK (10) —070.91

CK (11) —070. 91

CK ( 12) —070. 91

CK (13) —070. 91

CK (14) —09. 04

CK (13) —09. 04

C SURFACE EMITTANCE

DO 990 1-1,25

em ( I ) - .

9

S(I)«0.

990 CONTINUE

EM (7) —0.

EM (5) —0.

EM (9 ) -0.

EM ( 10) -0.

EM (12) —O.

EM (13) —0.

EM (24) —0.

C LIGHTING POWER

1 pm-2620.

C BALLAST FRACTION

BALF-O. 13

C SHORT WAVE FRACTIONS

a ( 1

) -O.

a ( 2 ) -O .

0

a (3) -0.

a (4) -O. 0296

S (6) -0.0734

S ( 7 ) —0 .

0

S (0) -. 4930

15) -0.0269.

127
uu

S ( 16) *0. 0923

S ( 17) “0. 096S

S ( 18) “0. 0939

S ( 19) =0. 0922

C LUMINOUS EFFICACY

tle-109.

C LIGHTING LUMENS

llm=51200.

c si gma- (0. 1714)*10.**(-8)

C TIME STEP ONE MINUTE

tao=l . /60.

C CONVECTION MULTIPLIER

HCM® 1.0

c TIME-0. -TAO

AF-200.

TADJ-O. 67

WRITE ( * , 500

)

500 FORMAT (' ENTER OUTPUT FILENAME (7 CHARS)

t

READ (*,505) FN

505 FORMAT (A7)

OPEN ( 7 f F I LE-FN , STATUS- ' NEW '

)

write(*,470)

470 -format (

' enter convection multiplier (1.)')

read (*,*)hcm

400 -f ormat (-f 10. 2)

C

do 10 ii-1,1450

C

LP-LPM* ( 1 . 3089-. 00323*T ( 4 )

)

LPL-1. 1 *LPM

IF (LP.GT.LPL) LP«LPM*1 u 1 LL—LLM* < 1 . 2960—. 00297*T (4)

)

LLL-1 . 1 *LLM

IF (LL.GT.LLL) LL=LLM*1.1

VW-LL/TLE

QBAL-BALF*LP

QSOR-LP—QBAL—VW

C

DO 20 1-4,25

qv ( i

) -s ( i ) *vw

C THERMAL RADIATION

eb < i ) -si gma* ( t ( i ) +460. ) **4

e-f ( i ) -em ( i ) / ( 1 . -em ( i ) ) C CONVECTION COEFFICIENTS

he (i ) -O. 22* (abs (t (i ) -t (1) ) **0. 33) *hcm*l .

1

20 continue

HC (5) —O.

HC (7) -0.

HC (9) —0.

HC (12) -0.

HC ( 4 ) —HC ( 4 ) *2

.

HC(6) -HC (6) *2.

DO 21 1=20,25

HC ( I

) -0.22* (ABS (T ( I

) -T (2) ) **0. 33) *HCM

21 CONTINUE

HC ( 1 1 ) -0. 22* (ABS (T ( 1 1

) -T (2) ) **0. 33) *HCM

HC ( 14) -0. 22* (ABS (T ( 14) -T (2) ) **0.33) *HCM

HC ( 24 ) —HC ( 24 ) *3

.

HC ( 25 ) —HC ( 25 ) *2

.

RADIATION HEAT BALANCE

do 30 i -1 , 30

MAXC-O.

128
nonnoonn

n

nn

n

n

nn

DO 35 J«4,25

SUM ( J ) *0.

DO 36 K=4 , 25

IF(K.EQ.J) GOTO 36

SUM ( J ) »SUM <J) + (F (J,K)*B(K)

)

36 CONTINUE

B(J)»(EF(J)*EB(J)+SUM(J) )/(l+EF(J)

)

CHANGE=ABS ( B ( J ) -OB < J )

)

IF < CHANGE. GT. MAXC) MAXC=CHANGE

OB < J ) = B ( J

)

35 CONTINUE

IF (MAXC. LT. .01) GOTO 38

30 continue

38 CONTINUE

WRITE (* ,507) 1 1 ,

1

507 FORMAT ( 1 X , 217)

IRRADIANCE ON NODES

DO 39 J-4,25

H(J)«0.

DO 39 K-4,25

IF(K.EQ.J) GOTO 39

H(J)«H(J)+(F(J,K)*B(K>

)

39 CONTINUE

do 40 i «4, 25

NET THERMAL RADIATION

qrn (i ) (h (i ) -b (i ) ) *a (i ) 40 continue

QRN (6) -QRN (6) +QRN (25)

QRN (5) *0.

QRN (7) *0.

QRN (9) =0.

QRN ( 10 ) =0.

QRN (12) “0.

QRN (13) *0.

QRN (24) *0.

CONVECTION TO ROOM AIR

qc(l)«hc (4) * (t (4) —t ( 1 )

) *a (4 ) +hc (6) * ( t (6) -t ( 1 > ) *a (6) * +HC ( 8 ) * ( T ( 8 ) -T ( 1 ) ) *A ( 8

)

* +HC ( 1 5 ) * ( T (15) -T ( 1 ) ) *A (15) +HC (16)*(T(16) —T ( 1 ) )*A(16)

* +HC (17)*(T(17> -T ( 1 ) > *A (17) t-HC ( 18) * (T ( 18) -T ( 1 ) )*A(18)

* +HC (19)*(T (19) —T ( 1 ) ) *A (19)

CONVECTION TO SURFACE NODES

do 50 i “4 ,19

qc(i)«tic(i)*(t(l)-t<i) )*a(i)

50 continue

QC (5) *CK (5) * (T (6) —T (5) ) QC(6)=QC(6)+CK(6)*(T(5)-T(6)

)

DO 55 1=20,25

QC ( I

) “HC ( I ) * ( T ( 2 ) -T ( I ) )*A(I)

55 CONTINUE

QC (11) »HC < 1 1 ) * ( T ( 2 ) -T (11) ) *A (11)

QC (14) “HC (14)*(T (2)-T (14) )*A(14)

WRITE (7, 949) (QC(J) , J»1 ,25)

949 FORMAT (IX ,25F10. 3)

CLD=QC (4) +QC (6) +QC (8) +QC (15) +QC (16) +QC (17) QC (18) +QC (19)

* - ( CK ( 6 ) * ( T ( 5 ) —T ( 6 ) )

)

CLU=QC (11) +QC ( 14) +QC (20) +QC (21 ) +QC (22) +QC (23) +QC (24) +QC (25)

CLT-CLD+CLU

WR I TE ( 7 , 948 ) CLD , CLU , CLT

948 RQRMAT (IX. 3F LO. 2)

129
on

on

QC (14) -QC (14) +CK (14)*(T ( 13) -T (14)

)

QC (IS) -QC (15) +CK ( 1 5 ) * ( T ( 1 4 ) -T (13) ) QC (8) -CK (8) * (T (9) —T (8) )+QC(8)

QC ( 9 ) —CK <9)*<T ( 8 ) -T (9) ) +CK (9) * (T ( 1 1

) -T (9)

)

QC ( 1 1

> -CK ( 1 1 ) * ( T ( 1 2 ) -T ( 1 1 ) » +QC (11)

QC (12) —CK ( 12) * (T (11) -T (12) ) +CK ( 12) * (T (8) -T ( 12)

)

QC (6) -QC (6) +QC (255)

C HEAT STORAGE

qs ( 1

) =qc ( 1 ) *tao

do 60 i-4,23

qs ( i ) * ( qrn ( i ) +qc ( i ) +qv ( i ) ) *t ao

60 continue

ADD LAMP SOURCE HEAT

qs (4) -qs (4) +qsor*tao

QS (3) -QBAL*TAO+QS (5)

QS (2) -TAO* (HC (23) * (T (23) —T (2) ) *A (25) +HC ( 1 1 ) * (T ( 1 1

) -T (2) ) *A ( 1 1

)

* +HC (14)*(T (14) —T ( 2 ) ) *A (14) +HC ( 20 ) * ( T ( 20 ) —T ( 2 ) )*A(20)

* +HC (21)*(T (21) —T ( 2 ) ) *A (21) +HC ( 22 ) * ( T ( 22 ) -T ( 2 ) )*A(22)

* +HC ( 23 ) * ( T ( 23 ) —T ( 2 ) ) *A (23) +HC (24) * (T (24) -T (2) )*A(24)

* — ( AF* e 0805*. 24*60*TADJ* (T (2) -T ( 1 ) ) )

)

Q4 1 -HC (4)*(T<4) —T ( 1 ) )*A(4)

Q61—HC (6) * (T (6) —T ( 1 ) )*AC6)

Q7 1 —HC ( 7 ) * ( T (7) -T ( 1 ) )*A(7)

Q62—HC (25) * (T (25) —T (2) >*A(23)

CL-QS ( 1 ) + (AF*60. *TADJ*. 0805*. 24* (T (2) -T ( 1 ) ) ) *TAQ

COMPUTE TEMPERATURE CHANGES

do 70 i-2,25

t(i)«t(i)+qs(i)/c<i>

70 continue

T (23) —T (6)

T ( 10) —T (11)

T ( 13) -T (8) c 1 f -cl / ( 1 p*TAQ)

TIME-TIME+TAO

048—A <4 ) *B (4 ) *F (4,8) —B (8 ) *A ( 8 ) *F (0,4)

Q4 16—B ( 4 ) *A (4) *F (4 , 16) —B ( 16) *A ( 16 ) *F (16,4)

Q4 17-B (4) *A (4) *F (4 , 17) —B (17) *A (17) *F (17,4)

Q418-B (4) *A (4) *F (4, 18) -B (19) *A (18) #F (18,4)

Q419«B(4)*A(4)*F(4, 19) -B ( 19) *A ( 19) *F < 19,4)

Q68—8 (6) *A (6) *F (6 , 8) —B (8) *A (8 ) *F (8,6)

Q6 16—B (6) *A (6) *F (6 , 16) -B ( 16) *A (16) *F (16,6)

Q617-B (6) *A (6) *F (6, 17) -B < 17 ) *A (17) *F (17,6)

Q610-B (6) *A (6) *F (6 , 18) —B ( 18) *A (IS) *F (18,6)

Q619»B(6)*A(6)*F(6, 19) -B < 19) *A ( 19) *F ( 19,6)

Q70-B ( 7 ) *A ( 7 ) *F ( 7 , 8 ) -8 < 8 ) *A ( 8 ) *F ( 8 , 7

)

Q716-B (7) *A (7) *F (7, 16) —B (16) *A (16) *F (16,7)

Q717-B (7) *A (7) *F (7 , 17) -B ( 17 > *A ( 17) *F ( 17 , 7)

Q718-B (7) *A ( 7 ) *F (7 , 10) —B ( 1@ ) *A (IQ) *F (18,7)

Q719-B(7)*A(7)*F(7, 19) -B ( 19) *A ( 19) *F ( 19 , 7)

Q61 i«B (25) *A (25) *F (25, 1 1

) -B ( 1 1 ) *A ( 1 1 > *F ( 1 1 ,23)

Q620-B (25) *A (25) *F (25,20) -B (20) *A (20) *F (20,25)

Q621-B (25) *A(25)*FC25,21)-B(21)*A(2U*FC2i ,25)

Q622-B ( 25 ) *A ( 23 ) *F ( 23 , 22 ) -B ( 22 ) *A ( 22 ) *F ( 22 , 25

)

Q623-B (25) *A (23) *F (23 ,23) -B (23) *A (23) *F (23,23)

I RD- ( Q48+Q4 1 6+Q4 1 7+Q4 1 8+Q4 1 9+Q68+Q6 1 6-M36 1 7+Q6 1 8+Q6 1 9 ) /L

P

IRU- (Q61 1+Q620+Q621+Q622+Q623) /LP

COD- (Q41+Q61 ) /LP

COU—Q62/LP

VD-VW/LP

VU-O.

COR-O.

STR— 1

. —VD—VU— I RD” I RU—COD—COU—COR

130
nnnnn

i 2=i i *1

IF < 1 1 . EQ. 0. OR. 1 1 . EQ. 1 . OR. II.EQ.10. OR. II. EQ. 30) GO TO 991

IF ( 1 1 . EQ. 30. OR. II. EQ. 60. OR. 1 1 . EQ. 90. OR. II. EQ. 120) GO TO 991

IF < 1 1 . EQ. 150. OR. II.EQ. 180. OR. 1 1 . EQ. 210. OR. 1 1 . EQ. 240) GO TO 991

IF < 1 1. EQ. 270. OR. II. EQ. 300. OR. II. EQ. 330. OR. II. EQ. 360) GO TO 991

IF < 1 1 . EQ. 390. OR. II. EQ. 420. OR. II. EQ. £50. OR. 1 1 . EQ. 480) GO TO 991

IF ( 1 1 . EQ. 510. OR. 1 1 . EQ. 540. OR. 1 1 . EQ. 570. OR . 1 1 . EQ. 600) GO TO 991

IF(II.EQ. 630. OR. II. EQ. 660. OR. 1 1 . EQ. 690. OR. 1 1 . EQ. 720) GO TO 991

IF ( 1 1 . EQ. 750. OR. II. EQ. 780. OR. II. EQ. 820. OR. II. EQ. BSO) GO TO 991

IF ( 1 1. EQ. 880. OR. 1 1 . EQ. 910. OR. 1 1 . EQ, 940. OR. 1 1 . EQ. 970) GOTO 991

XFdI.EQ. 1000. OR. II. EQ. 1030. OR. II.EQ. 1060. OR. II.EQ. 1090) GOTO 991

IFdI.EQ. 1120. OR. II.EQ. 1130. OR. II.EQ. 1 180. OR. II.EQ. 1210) GOTO 991

IF (II.EQ. 1240. OR. II.EQ. 1270. OR. II.EQ. 1300. OR. 1 1 . EQ. 1330) GOTO 991

IFdI.EQ. 1360. OR. II.EQ. 1390. OR. II.EQ. 1420. OR. II.EQ. 1430) GOTO 991

600 FORMAT ( 1 X , 1 3 , 25F 10.2)

GO TO 10

991 CONTINUE

WRITE (7,600) 12, (T(J) ,J=1 ,25)

WRITE (7,600) 12, (QS(J) , J-l ,25)

WRITE (7,600) 12, (QRN(J) , J-l, 25)

WRITE (7,600) 12, (QC(J) , J-l ,25)

WRITE (7,600) 12, (QV(J) ,J-1 ,25)

WR I TE ( 7 , 700 ) T I ME , CL , LP , LL , CLF , STR , BALF , VD , VU , IRD , I RU , COD , COR , COU

,

+ (T(IJ) ,IJ-1 ,25)

700 FORMAT (IX, 4F 10.2,1 0F7 . 4 , 23F6 . 1

)

10 continue

c CLOSE (7)

C

end

131
References

As77 American Society of Heating, Refrigerating and Air Conditioning

Engineers, Handbook of Fundamentals, 1979.

Ba83a Ball, H.D., Green, D., The Impact of Lighting Fixtures on

Heating and Cooling Loads - Mathematical Model, ASHRAE

Transactions Vol. 89, Pt 2A and B, 1983.

Ba83b Ball, H.D., The Impact of. Lighting Fixtures on Heating and

Cooling Loads - Application to Design, ASHRAE Transactions,

Vol. 89, Pt 2A and B, 1983.

Ba64 Ballman, T.L. , Bradley, R.D. , Hoelscher, E.C., Calorimetry of

Fluorescent Luminaires, Illuminating Engineering, Dec. 1964,

pg. 779-785.

Ba65 Ballman, T.L. Mueller, T. , Fluorescent Luminaire Calorimetry:

Principles and Procedures, ASHRAE Transactions Vol. 71, Pt 1,

1965, pg. 51-60.

Be77 Bedocs, L. , Godfrey, R.L., Sanford, T.P., Air-handling

Luminaires and Plenum Exhaust Systems, Lighting Research and

Technology, Vol. 9, No. 2, 1977, pg. 73-84.

Bo63 Bonvallet, G.G. , Method of Determining Energy Distribution

Characteristics of Fluorescent Luminaires, Illuminating

Engineering, Feb., 1963, pg. 69-74.

Bo77 Boyce, P.R., Air-Handling Luminaires, Lighting and

Environmental Design, Vol. 70, No. 6, Nov. 1977, pg. 243-276.

Bo66 Boyer, L.L., The Role of the Supply-Air Plenum in Interior zone

Recessed Illumination Systems, Illuminating Engineering, Aug.

1966, pg. 528-533.

Bo67 Boyer, L.L., Thermal Environment in Perimeter Office Spaces

Cooled with Ventilating Ceilings, ASHRAE Transactions, Vol. 71,

Pt 1, 1967.

Bo68a Boyer, L.L., Radiant Panel Effects of Floor-Ceiling Assemblies

Incorporating Static, Return and Supply Plenums, ASHRAE

Transactions, Vol. 74, Pt II, 1968, pg. 31-42.

Da40 Darley, W.G., Three in One-Light, Sound and Air Conditioning,

Magazine of Light, Vol. 9, No. 2, Feb. 1940, pg. 27-29.

Do79 DoE-2 Users Guide, Lawrence Berkeley Laboratory and Los Alamos

Scientific Laboratory, U.S. Dept, of Energy, Contract W-7405-

ENG-48 . Feb. 1979.

Du64a Dunn, G.W., Gilleard, G,. Schmidt, R.D., Advanced Electrical

Space Conditioning, Illuminating Engineering, Jan. 1964, pg.

43-51.

132
Du64b

E133

Fi76

Fi65

Fi66

Fi70

He41

He66

Hi79

Ie56

Ie66

Ie68

Ie72

Je73

Ki68

Dunn, G. W, , Gilleard, G., Schmidt, R.D., A Field Study of an

Application of the Heat Transfer Lighting Fixture, ASHRAE

Transactions, Vol. 70, 1964, pg. 347-353.

Electrical World, Removing heat from High-Intensity Lighting,

April 15, 1933, pg. 480-481.

Finn, J.F., Efficient Application of Lighting Energy-a

Luminaire Air Heat Transfer Evaluation, Lighting Design and

Application, Jan. 1976, pg. 40-43.

Fisher, W.S., Roehr, L. A., Smith, J.M., Some Factors Affecting

Heat Transfer in Luminaires, Illuminating Engineering, Jan.

1965, pg. 51-60.

Fisher, W.S., Stephens, W.R.

,

Skerkoski, E.C., Electrical Space

Conditioning in a Remodeled Office Building, Illuminating

Engineering, March 1916, pg. 150-168.

Fisher, W.S., Weinstein, S., Heat Transfer with Hegh Intens ity

Discharge Lamps and Luminaires, Illuminating Engineering, April

1970, pg. 185-192.

Heating, Piping and Air Conditioning, Lighting and Air

Conditioning Design Factors, Sept. 1941, pg. 605-608.

Henderson, T.A., Heat Transfer from Ceiling Plenums, Air

Condition, Heating and Ventilating, Nov. 1966, pg. 72-76.

Hittle, D. , The Building Loads Analysis and System

Thermodynamics (BLAST) Program, Version 2.0, Users Mannual Vol.

1 and Vol. 2, Technical Report E- 153/ADA072272 and ADA722730,

U.S. Army Construction Engineering Research Laboratory, June

1979.

Illuminating Engineering, Lighting and Air Conditioning, Sept.

1956, pg. 647-652.

Illuminating Engineering, Lighting and Air Conditioning, March

1966, pg. 123-147.

IES Transaction, Photometric and Thermal Testing of Air and

Liquid Cooled Heat Transfer Luminaires, Sept. 1968, pg. 485-

497.

IES Lighting Handbook, 5th Edition, Illuminating Engineering

Society of North America, 1972.

Jewess, W.

,

Pinniger, J.H., Thermal Effects in Lighting

Equipment, Light and Lighting, Vol. 66, No. 9, Sept. 1973.

Kimura, K. , Stephenson, P.G., Theoretical Study of Cooling Load

Caused by Lights, ASHRAE Transactions, Vol. 74, Pt II, 1968.

133
Lu38 Luckiesh, M. , Cooler Footcandles, Magazine of Light, Year End,

1938, pg. 23-38.

Ma71 Mathews, C.P., Ceiling Temperature Equations, Building Systems

Design, April, 1971, pg. 37-38.

Ma84 Maya, J., Grossman, M.W.

,

Lagushenko, R. , Waymouth, J.F.,

Energy Conservation Through More Efficient Lighting, Science,

Oct. 26, 1984, pg. 435-436.

Mc76 McIntyre, D.A.

,

Radiant Heat from Lights and its Effect on

Thermal Comfort, Lighting Research and Technology, Vo 1. 8, No.

3, 1976, pg. 121-128.

Mc83 McKinley, A.D.

,

Mitalas , G.P., Room Thermal Transfer Functions,

- Computer Programs to Calculate z -Transfer Function

Coefficients for Rooms, Computer Program No. 52, National

Research Council Canada, De . 1983.

Me63 Meckler, G. , Meckler, M. , Design and Evaluation of Dynamically

Integrated Lighting-Air Conditioning Systems, Illuminating

Engineering, Feb. 1963, pg. 75-85.

Me64 Meckler, G., Pennington, C.W., Humphreville , T.N., Energy

Integrated Lighting-Heating-Cooling System, ASHRAE

Transactions, Vol. 70, 1964, pg. 335-346.

Mi71 Mitalas, G.P., Kimura, K. , A Calorimeter to Determine Cooling

Load Caused by Lights, ASHRAE Transactions, Vol. 77, Pt II,

1971, pg. 65-72.

Mi72 Mitalas, G.P., Transfer Function Method of Calculating Cooling

Loads, Heat Extraction and Space Temperature, ASHRAE Journal,

Vol. 14, No. 12, Dec. 1972 pg. 54-56.

Mi73 Mitalas, G.P., Calculating Cooling Load Caused by Lights,

ASHRAE Journal, Vol. 15, No. 6, June 1973. pg. 37-40.

Mi74 Mitalas, G.P., Cooling Load Caused by Lights, Transactions

Canadian Society for Mechanical Engineering, Vol. 2, No. 3,

1973-74, pg. 169-174.

Mi78 Mitalas, G.P., Comments on the z -Transfer Function Method for

Calculating Heat Transfer in Buildings, ASHRAE Transactions,

Vol. 84, Pt 1, 1978, pg. 667-674.

Mi83 Mitalas, G.P., Room Dynamic Thermal Response, DBR Paper No.

1189, National Research Council Canada, 1983.

Mo40 Mortensen, M.A.

,

Continuous Trough Lighting, Magazine of Light,

Vol. 9, No. 2, Feb. 25, 1940, pg. 30.

Mu62 Mueller, T., Benson, B.Sg„, Testing and Performance of HeatRemoval Troffers, Illuminating Engineering, Dec. 1962, pg. 793-

802.

134
Mu85 Murdoch, J.B., Illuminating Engineering, Macmillan Publishing

Co., New York, 1985.

Ne71 Nevins , R.G.

,

Straub, H.E., Ball, H.D., Thermal Analysis of

Heat Removal Troffers, ASHRAE Transactions, Vol. 77, Pt II,

1971, pg. 58-64.

No69 Nottage, H.B.

,

Park, K.S., Performance of a Ventilated

Luminaire, ASHRAE Transactions, Vol. 75, Pt I , 1969, pg. 50-70.

0b70 O'Brien, P.F., Krumland, L.R.

,

An Investigation of the Luminous

and Thermal Radiative Transfer Characteristics of Luminaires,

Illuminating Engineering, April 1970, pg. 218-230.

Qu62 Quin, M.L.

,

Bradley, R.D.

,

Corrective Transfer of Lighting Heat

Loads, Illuminating Engineering, Jan. 1962, 45-56.

Ro78 Raven, F.H.

,

Automatic Control Engineering, McGraw-Hill Book

Co., 1978.

Ro76 Rowe, G.D., Selecting Recessed Fluorescent Troffers for Optimum

Efficiency and Visual Comfort, Lighting Design and Application,

Feb. 1976, pg. 18-23.

5184 Siminovitch, M.J., Rubinstein, F.M. Verderber, R.R.

,

Determining Lamp/Ballast System Performance with a TemperatureControlled Integrating Chamber, Journal of IES , Oct. 1984, pg.

364-378.

5185 Siminovitch, M.J., Rubinstein, F.M.

,

Clark, T.A., Verderber,

R.R.

,

A Luminaire/Plenum/HVAC Simulator, IEEE-IAS Conf. Paper,

Oct. 1985.

Si87 Siminovitch, M.J., Rubinstein, F.M.

,

Clark, T.A., Verderber,

R.R.

,

Maintaining Optimum Fluorescent Lamp Performance Under

Elevated Temperature Conditions, Lighting Design and

Application, Sept. 1987, pg. 35-41.

5072 Sowell, E.F., Environmental Radiation from Fluorescent Ceiling

Systems, Ph.D. Dissertation, University of California, Los

Angeles, 1972.

5073 Sowell, E.F., O'Brien, P.F., The Transport of Lighting Energy,

ASHRAE Transactions, Vol. 79, Pt II, 1973, pg. 172-182.

5074 Sowell, E.F., A General Model for the Transport of Lighting

Thermal Power, Second Symposium on the Use of Computers for

Environmental Engineering, Paris, France, 1974, pg. 85-113.

Sp84 Spitler, J.D., An Experimental Study of Interior Convective

heat Transfer in Buildings, Masters Thesis, University of

Illinois at Urbana- Champaign, 1984.

135
St69 Steck, B. , European Practice in the Integration of Lighting,

Air Conditioning and Acoustics in Offices, Lighting Research

and Technology, Voi. I, No. 1, 1969.

Sy64 Sylvester, G.C., Performance of Heat-Removal Troffer with a

Specific Building Design, ASHRAE Transactions, Vol. 70, 1964,

pg. 328-334.

Un85 United States Department of Energy, Overview of Building Energy

Use and Report of Analysis - 1985, Office of Building and

Community Systems, Wash. DC, Oct. 1985.

Wa86 Walton, G.N.

,

Algorithms for Calculating Radiation View Factors

Between Plane Cover Polygons with Obstructions, NBSIR 86-3563,

National Bureau of Standards, Oct. 1986.

136
NBS-1MA (REV. 2-«C) U.S. DEPT. OF COMM.

BIBLIOGRAPHIC DATA

1. PUBLICATION OR

REPORT NO.

2. Performing Organ. Report No. 3. Publication Date

SHEET (See instructions

)

NISTIR 88-3896 DECEMBER 1988

4. TITLE ANO SUBTITLE

The Interaction of Lighting, Heating and Cooling S ^tems in Buildings-Interim Report

5. AUTHOR(S) --r

Sr p.phen J. Treado

6. PERFORMING ORGANIZATION (If joint or other than NBS. see instructions) 7. Contract/Grant No.

NATIONAL BUREAU OF STANDARDS

8. Type of Report & Period Covered

U.S. DEPARTMENT OF COMMERCE

GAITHERSBURG, MD 20899

9. SPONSORING ORGANIZATION NAME AND COMPLETE ADDRESS (Street. City. State. ZIP)

US DOE

10. supplementary notes

j

I 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. I f document includes a si gnificant

bi bliography or literature survey, mention it here)

The effect of interactions between building lighting, heating and cooling systems on the

energy performance of the lighting system and cooling loads is examined, based on

detailed full-scale measurements and supporting computer simulations. A test facility

was designed, constructed and operated to emulate an office space with recessed fluorescent lighting. Special design features simulated adjacent building areas, providing

realistic thermal boundary conditions for the test room. The test facility was extensively instrumented to monitor lighting power, cooling load, surface and air temperatures

,

heat flows and light levels. 398 measured parameters were averaged and recorded every

two minutes during testing.

This interim report describes preliminary results from the research effort. The results

showed that the lighting system can be constrained to operate at its most efficient

level, if the fluorescent lamps are cooled sufficiently.

A two-term exponential relation with four regression coefficients was found to fit the

measured data well. An extended set of weighting factors was derived based on the

double exponential relation, and values computed from the measurements. Cooling load

profiles due to cyclic operation of the lights were computed using the weighting factors

Lighting energy distribution fractions were determined using a calibrated computer modeL

Design considerations are discussed.

12. KEY WORDS (Six to twelve entries; alphabetical order; capitalize only proper names; and separate key

building energy, cooling, heat transfer, lighting, peakloads, weighting

words by semicolon s) factors

13. AVAILABILITY

[X i U n 1 i mi ted

j

i For Official Distribution. Do Not Release to NTIS

1 Order From Superintendent of Documents, U.S. Government Printing Office, Washington, D.C.

20402. •

14. NO. OF

PRINTED PAGES

152

15. Price

/ Order From National Technical Information Service (NTIS), Springfield, VA. 22161

$ 18.95

USCOMM-OC S043-P80
'