4

- i CHAPTER 4

. ENERGY USE CHARACTERIZATION

TABLE OF CONTENTS

4.1 INTRODUCTION

................................ ................................ ................................ .............. 4

- 1 4.2 ENERGY USE ANALYSIS BY EQUIPMENT TYPE

................................ ..................... 4

- 1 4.2.1 Water

- Cooled Air Conditioners

................................ ................................ .............. 4

- 1 4.2.1.1 Water

- Cooled Compressor

- Condenser Performance

................................ .. 4

- 3 4.2.1.2 Supply Water Temperature

................................ ................................ ......... 4

- 5 4.2.1.3 Supply Fan Power

................................ ................................ ....................... 4

- 5 4.2.1.4 Water

- Cooled Compressor

- Condenser

- Only Cooling Coefficient of Performance vs. Energy Efficiency Ratio Relationship

................................ ......... 4

- 6 4.2.1.5 Baseline Annual Energy Use Per Ton

– Water

- Cooled

.............................. 4

- 7 4.2.1.6 Energy Use Estimates for Baseline and Higher Efficiency Water

- Cooled Air Conditioner Equipment by Equipment Class.

................................ ......................... 4

- 9 4.2.2 Evaporatively Cooled Air Conditioners

................................ ................................ 4

- 11 4.2.2.1 Baseline Annual Energy Use Per Ton

– Evaporatively Cooled

............... 4

- 12 4.2.2.2 Energy Use Estimates for Baseline and Higher Efficiency Evaporatively Cooled Air Conditioners by Equipment Class

................................ ...................... 4

- 13 4.2.3 Computer Room Air Conditioners

................................ ................................ ........ 4

- 14 4.2.3.1 Development of Computer Room Air Conditioner Performance Map

..... 4

- 15 4.2.3.2 Estimate the Average Computer Server Heat Load

................................ .. 4

- 15 4.2.3.3 Estimate the Condensing Fluid or Air Temperature for the Computer Room Air Conditi oner Unit

................................ ................................ ............................. 4

- 16 4.2.3.4 Economizer Load Reduction

................................ ................................ ..... 4

- 18 4.2.3.5 Estimate the Cooling Load

................................ ................................ ........ 4

- 20 4.2.3.6 Estimate the Hourly Power Usage

................................ ............................ 4

- 20 4.2.3.7 Calculate Annual Energy Consumption by Climate and by State

............ 4

- 24 4.2.4 Water

- Source Variable Refrigerant Flow Heat Pumps

................................ ......... 4

- 24 4.2.4.1 Building Model Characteristics

................................ ................................ 4

- 24 4.2.4.2 Variable Refrigerant Flow Modeling Characteris tics

............................... 4

- 28 REFERENCES

................................ ................................ ................................ .......................... 4

- 32

LIST OF TABLES Table 4.2.1 Normalized Energy Use per Ton by Climate for Water

- Cooled Air Conditioners, 11.0 Energy Efficiency Ratio

................................ ................................ ............... 4

- 9 Table 4.2.2 Water

- Cooled Air Conditioner Representative Capacities by Equipment Class

.... 4

- 10 Table 4.2.3 Water

- Cooled Air Conditioner Unit Energy Consumption by Equipment Class

... 4

- 11 Table 4.2.4 Normalized Energy Use Ton by Climate for Evaporatively Cooled Air Conditioners, 11.0 Energy Efficiency Ratio

................................ ................................ ............. 4

- 13 Table 4.2.5 Evaporatively Cooled Air Conditioner Representative Capacities by Equipment Class

................................ ................................ ................................ ................... 4

- 13 Table 4.2.6 Evaporatively Cooled Air Conditioner Unit Energy Consumption

........................ 4

- 14 Table 4.2.7 Condensing Conditions Established for CRAC Equipment

................................ ... 4

- 184

- ii Table 4.2.8 CRAC Performance as a Function of Entering Condenser Fluid or Outdoor Air Temperature

................................ ................................ ................................ ....... 4

- 22 Table 4.2.9 Envelope Performance Characteristics

................................ ................................ ... 4

- 25 Table 4.2.10 VRF Modeling Characteristics

................................ ................................ ............. 4

- 29 Table 4.2.11 Climate and Locations for Simulations

................................ ................................ 4

- 29 Table 4.2.12 National Average VRF Energy Use

– Simulations Results

................................ .. 4

- 30 Table 4.2.13 Unit Energy Consumption Estimates by Efficiency Level for Water

- Source VRF Equipment

................................ ................................ ................................ .......... 4

- 31

LIST OF FIGURES Figure 4.2.1 Comparison of Floor

- Space Weighting Factors by Region, Small and Medium Office Buildings versus All Buildings

................................ ................................ . 4

- 8 Figure 4.2.2 Cooling EIR as a Function of Entering Condenser Fluid Air Temperature

.......... 4

- 22 Figure 4.2.3 Small Office Building Design View

................................ ................................ ...... 4

- 25 Figure 4.2.4 Small Office Lighting Schedule

................................ ................................ ............ 4

- 26 Figure 4.2.5 Small Office Plug Load Schedule

................................ ................................ ......... 4

- 27 Figure 4.2.6 Small Office Occupancy Schedule

................................ ................................ ........ 4

- 27 Figure 4.2.7 Small Office Thermostat Set Point Schedule

................................ ........................ 4

- 284

- 1 CHAPTER 4. ENERGY USE CHARACTER IZATION

4.1 INTRODUCTION
The n ational e nergy s avings characterization described in this chapter provides estimates of the energy savings consumers would realize from the establishment of s tandards at the levels set forth in American Society of Heating, Refrigerating and Air

- Conditioning Engineers ( ASHRAE ) Standard 90.1

- 2010 for specific classes of equipment analyzed as well as for m ore energy efficient s tandards for the same equipment classes

. This chapter des cribes the energy use analysis for water

- cooled and evaporatively cooled commercial packaged air conditioners , computer room air condition ing (CRAC) equipment , and water

- source variable refrigerant flow (VRF) equipment with a cooling capacity greater than 135,000 British thermal units per hour (B tu/h )

. These equipment types are described in separate subsections of this chapter , which detail the analysis and the determin ation of baseline annual per unit ene rgy consumption (UEC) estimates for specific equipment classes within each of the three equipment types examined

. Th is chapt er also describes the development of UEC estimates for
more
- efficient equipment, up to the maximum technology feasible ( max
- tech ) levels defined in either chapter 3 of the technical support document (TSD) or this chapter

. For each of these equipment types, the existing or proposed Federal standard is expressed
as an efficiency metric or metrics : e nergy e fficiency r atio ( EER ) for cooling efficiency , c oef f icient of p erformance (COP) for heating efficiency , or sensible coefficient of performance (SCOP) for the cooling efficiency of computer room air conditioners

. For each equipment class, this chapter describes how an estimate of the UEC is developed corresponding to different rated efficiencies

. 4.2 ENERGY USE ANALYSIS BY EQUIPMENT TYPE

4.2.1 Water

- Cooled Air Conditioners The analysis to assess the per unit energy consumption of water

- cooled air conditioners began with a review of the existing market as well as the review of historical shipments data provided by the Air

- Conditioning, Heating, and Refrigeration Institute ( AHRI ) and discussed in
c hapter 8 of the TSD

. The review s uggested that most water

- cooled air conditioner units on the market are designed for installation inside commercial buildings (as opposed to on building rooftops) . T he shipments data suggested that , in recent years , shipments were dominated by larger equip ment ( greater than 240,000 Btu/h capacity), with a trend toward relatively few shipments of smaller capacity units. For that reason , the U.S. Department of Energy ( DOE ) analysis of energy savings focused on typical applications for equipment in the 240,000 Btu/h or greater class

. M anufacturer literature suggested that a common application is floor

- by

- floor cooling in a multi

- story building. 1
To estimate the energy use of water

- cooled air conditioners in this application, DOE used annual hourly simulation d ata developed from computer simulations of a prototypical , medium sized, commercial office building

. The prototype building model was a three

- story, 53,600 ft 24

- 2 commercial office building developed as part of DOE’s commercial reference building models

. a In t his building model , referred to as a “ medium office, ” each floor is served by a separate packaged air conditionin g unit. The hourly data used for this analysis was previously developed from simulations using this model and the DOE EnergyPlus building simu lation software , and reflected building simulations in 15 clim ate locations in the United States

. b E ach climate location used is representative of a climate in 1 of 15 climate regions that have been developed in DOE’s Building Energy Codes Program and subs equently used in the development of the commercial reference building models

. 2 , 3 Characteristics of the DOE reference medium office building model used in this analysis are provided in A ppendix 4 A

. The medium office r eference building model u sed packaged variable air volume (VAV) rooftop cooling units , one unit serving each floor of the building, and the hourly thermal load and supply fan air volume data used for this analysis reflects this. DOE determined that the hourly cooling thermal loads fr om modeling of this type of equipment would be representative of hourly loads for systems served by larger water

- cooled equipment also providing floor

- by

- floor cooling and serving multiple building thermal zones with a VAV air distribution network . Thus, t hese loads would be accurate for the analysis of water

- cooled air conditioners in this application. EnergyPlus does not have an equipment component model developed around a water

- cooled air conditioner used in this application

. 4 EnergyPlus does have equipm ent models of water

- source heat pumps, but these equipment models do not appear to provide for variable volume airflow

. For this reason , DOE used the previously developed hourly cooling thermal load, hourly airflow , and hourly air temperature data for the air

- cooled packaged rooftop equipment used in the medium office reference building model as a starting point for its analysis.

To convert the hourly data to annual equipment energy consumption for water

- cooled air conditioners for baseline

- level equipment , DOE developed a spreadsheet model of the equipment performance of a water

- cooled air conditioner using actual manufacturer performance data for a n existing 25

- ton water

- cooled air conditioner

. 5 Cooling capacity and combined compressor

condenser c power consumption curve fits to this data were developed using polynomial relationships and the independent variables recommended for modeling of cooling efficiency for water

- source heat pumps in EnergyPlus ( with minor variations )

. Discussion of the model ing of the water

- cooled condenser and supply fan power follow s

.

a The commercial reference building models are available on DOE’s Office of Energy Efficiency and Renewable Energy website as Energy Plus input files at:
< http://www1.eere.energy.gov/buildings/commercial_initiative/new_construction.html >. Documentation of the model development is provided in Deru, M . et al. , U.S. Department of Energy Commercial Reference Building Models of the National Building Stock ( NREL/TP

- 5500

- 46861 ) (2011)

. b Hourly simulation results for the medium office reference building model were developed at Pacific Northwest National Laboratory (P NNL) and provided for support of the ASHRAE 90.1 Mechanical Subcommittee in April 2009. Simulation results used reflect code requirements for buildings designed to ANSI/ASHRAE/IESNA 90.1

- 2004.

c Compressor

- condenser refers to the combined power (or effic iency) of the compressor and condenser part of the equipment. It does not reflect the integration of the supply fan power consumption. This terminology is used throughout the energy analysis characterization discussion.4

- 3 4.2.1.1 Water

- Cooled Compressor

- Condenser P erformance EnergyPlus provides two methods of modeling equipment performance of water

- source heat pumps. The method used for this work is referred to as the “ eq uation

- fit model. ” The equation

- fit model uses non

- dimensional equations developed from curve

- fits to manufacturer’s performance data . These curve fits are used to predict the condenser performance in cooling (and heating ) mode

. L east square s regression is used to generate a set of performance coefficients from catalog performance data at various conditions for the equipment modeled

. Then , the respective coefficients from the least squares regressions are used in the model to simulate the heat pump performance in the EnergyPlus software

. The variables used by EnergyPlus to describe the water

- cooled compressor
- condenser performance are the entering air wet bulb and dry bulb temperature, the condenser water inlet temperature, the entering air volumetri c flow rate, and the condenser water flow rate. Each of these variables is made non

- dimensional by first dividing by corresponding values for each variable at a defined reference condition. The same method was used for the spreadsheet model. Two minor vari ations from the implementation in EnergyPlus included the use of i nch

- p ound units in developing the DOE performance curves and the specific selection of reference temperature conditions that correspond precisely to the DOE rating conditions. The governing equations for the cooling performance are as follows:

Q total Q total

_ ref = A 1 + A 2 × � T wb T

wb

_ ref � + A 3 × � T w , in T

w , in

_ ref � + A 4 × � V ̇ air V ̇ air

_ ref � + A 5 × � V ̇ w V ̇ w

_ ref �
Eq. 4

. 1

Q sen Q sens

_ ref = B 1 + B 2 × � T db T

db

_ ref � + B 3 × � T wb T

wb

_ ref � + B 4 × � T w , in T

w , in

_ ref � + B 5 × � V ̇ air V ̇ air

_ ref � + B 6 × � V ̇ w V ̇ w

_ ref �

Eq. 4

. 2

Power Power ref = C 1 + C 2 × � T wb T

wb

_ re f � + C 3 × � T w , in T

w , in

_ ref � + C 4 × � V ̇ air V ̇ air

_ ref � + C 5 × � V ̇ w V ̇ w

_ ref �
Eq. 4

. 3

Where :
A 1 through A 5 = e quation fit coefficients for the total cooling capacity , B 1 through B 6 = e quation fit coefficients for the sensible cooling capacity, C 1 through C 5 = e quation fit coefficients for the power consumption, Q total = c ooling c apacity (Btu/h) , Q total_ref = c ooling capacity at r eference c ondition (Btu/h) , Q

sens = s ensible cooling capacity (Btu/h) , Q

sens_ref = s ensible cooling capacity at reference condition (Btu/h) , Power = compressor
- c ondenser power (W) , Power ref = compressor
- c ondenser power at reference condition (W) , T

wb = e ntering air wet bulb t emperature ( Rankine ( ° R) ) , T

wb

_ ref = e ntering air wet bulb temperature at reference temperature ( ° R) , T

d

b = e ntering air dry bulb temperature ( ° R) , T

d

b

_ ref = e ntering air dry bulb temperature at reference temperature ( ° R) ,4

- 4 T

w ,in = e ntering water temperature ( ° R) , T

w ,in_ ref = e ntering water temperature at reference temperature ( ° R) , V ̇ air = v olumetric airflow rate ( cubic feet per minute ( cfm ) ) , V ̇ air

_ ref = v olumetric airflow rate at reference ( cfm ) , V ̇ w = v olumetric water flow rate ( gallons per minute ( gp m ) ) , and V ̇ w

_ ref = v olumetric water flow rate at reference (gpm)

. The equations above are based on steady state full

- load operation at the rating (reference) conditions and off

- rating conditions. To account for operation at part

- load and non
- steady

- state operation, part

- load performanc e modifier curves for air

- source air conditioners

— which were used in the EnergyPlus medium office reference building model

— were also used for this analysis

. Because these part

- load modifier curves reflect the effects of compressor cycling at part

- load for air

- cooled systems , it was determined that the y should be representative of the compressor cycling impacts for water

- cooled air conditioners as well.
The part

- load modifier curves are of the following equation form :
Cool

- PLF

- FPLR = C 1 + C 2 × PLR Eq. 4

. 4

Where:

Cool

- PLF

- FPLR = c ooling power part

- load factor modifier , PLR = r atio of cooling load to cooling capacity under the conditions of interest , and C 1 and C 2 = e quation fit coefficients for the part

- load performance curves

. C 1 and C 2 values of 0.771 and 0.229 , respectively , were used based on the default part

load modifier curves used in the reference building simulations and for this analysis

. The steady state power input is then adjusted according to the following equation :

Power Hourly = Power × PLR Cool

- PLF

- FPLR Eq. 4

. 5

Where :

Power Hourly = h ourly average power in response to hourly cool ing load , Power = s teady s tate power at conditions as defined in Eq. 4.3 , PLR = p art

- l oad r atio as defined in Eq. 4.4 , and Cool

- PLF

- FPLR = p art

- l oad modifier as defined in Eq. 4.4

. F or each climate zone , the original system air volumes were sized based on design day runs and the maximum hourly supply air for each system was used as the basis for the system air volumes and reference volumetric flow rate

. The peak cooling loads were adjusted for rated conditions using Eq. 4.1 and used to determine the equipment nominal capacity for each system. This is also the cooling capacity defined at the reference condition. No hourly condenser

- water4

- 5 flow rates were available , and for the modeling purpose t he condenser water flow rate at all hours was assumed to be at the reference (rated ) condition

. Thus , the condenser water flow rate did not affect the performance in the analysis

. 4.2.1.2 Supply Water T emperature The performance equations developed in this spread sheet model separately account for the water

- cooled gross cooling capacity and power consumption as a function of entering air conditions and supply water temperature and flow rate. To function, the spreadsheet model requires an hourly entering water tempe rature and entering water flow rate. For this analysis , a simple cooling tower supply water temperature model was developed based on a defined control profile with minimum 70 °F return water temperature and using a 7 °F approach temperature. Thus, the return water temperature from the cooling tower is calculated as

T

w = minimum � T wb , o + 7 , 70 � + 459

. 67 Eq. 4

. 6

Where :

T

w = b uilding s upply w ater t emperature from c ooling t ower ( °R ), and T

wb,o = o utdoor air wet bulb temperature ( °F ). The supply water temperature from the cooling tower is thus estimated at 7 °F above the current wet bulb temperature during most cooling hours

. Condenser water flow rates were assumed to be equivalent to the nominal rating c ondition water flow rates for all cooling hours. Using these inputs and performance curves , it was possible to calculate the hourly compressor
- condenser energy consumption of each of the three water

- cooled air cond itioner units for this building model.

4.2.1.3 Su pply Fan Power In addition, DOE’s spreadsheet model calculated the fan energy consumption for the equipment. Fan power for each system at design conditions (peak fan capacity) was estimated using the following assumptions :

Power fan , max = ∆ P × Q 6350 × n f × n m × 0

. 746 Eq. 4

. 7

Where :

Power fan,max = d esign f an p ower (kW) , Δ

P = t otal f an s tatic p ressure (in. H 2 0) , Q = d esign s upply airflow r ate ( cfm ) , n

f = f an e fficiency , and
n m = m otor and d rive e fficiency

.4

- 6 A design total fan static pressure of 4.35 in. H 2 O, f an efficiency of 0.65, and m otor
e fficiency of 0.9 were used, providing for a design fan system power of 0.87 W/cfm at design conditions. The total fan static pressure was judged likely to be less than that used in the reference building design (4.35 in. H 2 O versus 5.35 in. H 2 O) based on the determination that , due to the length of the duct runs, the total supply and return path pressure drop would be less with water

- cooled units installed insi de the building as opposed to packaged units installed on rooftops.

Because this was a VAV system with varying hourly airflow rate, the fan power was adjuste d hourly using a fan power part

- load curve used for the DOE reference building VAV system. The fan power curve provides the relationship between fan system power and system airflow rate, which varied hourly for the VAV system. The fan power versus flow rate relationship is of the form used in EnergyPlus:

PLF fan = C 1 + C 2 × FF + C 3 × FF 2 + C 4 × FF 3 + C 5 × FF 4 Eq. 4

. 8

Where :

PLF fan = a part

- load factor multiplier equal to the ratio of system fan power to design fan power , FF = f raction of system de sign airflow rate , and C 1 through C 5 = e quation coefficients for the VAV fan power performance curve

. The coefficients C 1 through C 5 used for th e VAV fan power performance curve are those from the original packaged VAV system used in the reference building model and are as follows:

C 1 = 0.040759894 C

2 = 0.08804497 C

3 = - 0.07292612 C

4 = 0.943739823 C

5 = 0

4.2.1.4 Water

- Cooled Compressor

- Condenser

- Only Cooling Coefficient of Performance vs . Energy Efficiency Ratio R elationship For analysis of energy use at the base efficiency (EER) level, DOE devel op ed estimates for the compressor
- con denser

- only efficiency ( compressor
- c ondenser

- only c ooling COP ) for the water

- cooled air conditioner based on the nom inal rating conditions for an 11 EER unit

. Compressor

- c ondenser

- only cooling COP at rating conditions is an input to both the EnergyPlus simulation and DOE’s spreadsheet model s and is the ratio of the gross output divided by the compressor
- condenser power, in consistent units

. This param eter can be calculated by backin g out the estimated fan power at nominal rating conditions from the rated energy consumption while accounting for the effect of fan heat

. This is do ne by calculating the compressor
condenser

- only cooling COP according the following equation:4

- 7 COP CU

_ C ool = EER 3

. 413 + R 1

− R Eq. 4

. 9

Where:

EER = EER rating of the equipment being modeled , R = ratio of supply fan power to total equipment power at the AHRI rating condition, and COP CU

_ Cool = Compressor

- condensing

- only cooling COP at rated conditions. The parameter R can be estimated from available manufacturer fan power data at the external static rating condition and a known airflow rate. Generally , however , the airflow rate at the rat ed condition is not provided in the manufacturer’s literature. DOE used an R of 0. 164 based on review of available data for a large water

- cooled air conditioner and assuming a supply airflow rate of 350 cfm/ton. 5 For an 11 .0 EER unit that DOE modeled directly, the calcula ted condenser

- only COP corresponding to an 11 .0 EER unit was 4.052

. 4.2.1.5 Baseline Annual Energy Use Per Ton

– Water

- Cooled
For each of the 15 climates locations , DOE used the spreadsheet model to calculate the hourly compressor
- condenser energy use and supply f an energy use for each of the three water

cooled air conditioner units in the r eference medium office building model and summed the results to provide annual energy use estimates for compressor
- condenser and for supply fan separately

. DOE also calculated the cooling capacity for each unit and summed these for the whole building

. Dividing the energy use figures for both the compressor
- condenser and the supply fan by the rated cooling capacity provides a normalized energy

- use
- per
- ton figure for each end use at the baseline efficiency level. These climate

- specific results , along with the regional weighting factors correspond to each climate , are shown in Table 4.2

. 1

. The climate weighting factors are developed from construction statistics for commercial office floor space for small and for medium sized office buildings (likely to use commercial packaged air conditioning equipment regardless of system size) for the period from 2003 to 2007

. 6 For each building type , weights were first calculated for the building class and then averaged. Figure 4.2

. 1 shows the floor space weighting factors for th e office buildings represented in this analysis compared with that of all 1 5 DOE reference buildings types with construction floor space tallied in the study.4

- 8 Figure 4.2

. 1 Comparison o f Floor

- Space Weighting Factors b y Region , Small and Medium Office Buildings v ersus All Buildings In addition, Table 4.2

. 1 shows a parameter called the “f ull l oad e quivalent o perating h our s ” (FLEOH), which is c alculated as the annual cooling load served, divided by the system cooling capacity. In this case , the F LEOH shown is the result from weighting of the FLEOH calculated for each individual air conditioning unit by the relative capacity of th at unit to the total capacity of all units serving the reference building model. While not used in DOE’s energy calculat ion, FLEOH values provide a useful reference point to compare with other energy calculation methods. 0% 5% 10% 15% 20% 25%

1A 2A 2B 3A 3B 3C 4A 4B 4C 5A 5B 6A 6B 7 8Relative Fraction of Construction by Climate Region

DOE Climate Region All Buildings Office Buildings4

- 9 Table 4.2

. 1 Normalized Energy Use p er Ton by Climate for Water

- Cooled Air Conditioners, 11 .0 E nergy E fficiency R atio

Climate

Zone Representative Climate Location Construction Wt. Annual Compres s or Condenser

Energy kWh/ton Annual Fan Energy kWh/ton Annual System Energy kWh/ t on FLEOH hours

1A Miami, F L 1.82% 1 , 999 257 2 , 256 2 , 718 2A Houston, TX 16.20% 1 , 440 215 1 , 654 1 , 966 2B Phoenix, AZ 4.99% 1 , 446 354 1 , 800 2 , 052 3A Memphis, TN 14.91% 1 , 051 204 1 , 255 1 , 440 3B El Paso, TX 10.15% 973 284 1 , 258 1 , 413 3C San Francisco, CA 1.82% 497 273 770 736 4A Baltimore, MD 18.18% 824 216 1 , 040 1 , 135 4B Albuquerque , NM 0.72% 776 357 1 , 133 1 , 154 4C Salem, OR 2.71% 517 247 764 736 5A Chicago, IL 16.96% 674 215 889 926 5B Boise, ID 5.70% 542 283 825 807 6A Burlington, VT 0.56% 545 230 775 759 6B Helena, MT 4.62% 446 360 807 682 7 Duluth, MN 0.56% 352 249 600 496 8 Fairbanks, AK 0.10% 259 436 696 398 Nat ional 100.00% 947 242 1 , 189 1 , 319 kWh = kilowatt

- hours
4.2.1.6 Energy Use Estimates for Baseline and Higher Efficiency Water

- Cooled Air Conditioner E quipment by Equipment C lass. National average estimates of the e nergy use for efficiencies other than th ose directly modeled in the spreadsheet tool were developed by first scaling the national average compressor
condenser energy use per ton from the spreadsheet simulation of 11.0 EER equipment ( shown in Table 4.2

. 1 ) by the condenser

- only cooling COP corresponding to the higher EER equipment

. The condenser

- only cooling COP is developed based on Eq. 4.9 for higher EER values. To that result is added the baseline fan energy use per ton values, and the sum is then multiplied by the selected representative tonnage for each equipment class ; or in equation form ,

UEC EER = CAP × [ CondUEC n × COP CU

_ Cool

_ Modeled COP CU

_ Cool

_ EER + FanUEC n ] Eq. 4

. 10

Where :
CAP = n ominal equipment capacity in cooling tons , UEC EER = unit energy c onsumption at the specific EER level being evaluated , CondUEC n = n ormalized c ondenser energy cons umption from the modeling, kWh/t on

– n ational basis , COP CU_Cool_M odeled = the c ooling compressor
- condenser

- only COP used in the spreadsheet equipment model ,4

- 10 COP CU

_ Cool

_ E

ER = t he cooling compressor
- condenser

- only COP corresponding to the specific EER level being evaluated , and FanUEC n = n ormalized compressor
- c ondenser energy consumption from the modeling, kWh/ t on

– n ational

. Thus, for example, to estimate the annual energy consumption of a 35

- ton water

- cooled air conditioner at an EER of 12 .4 , the initial step is to calculate a compressor
- condenser

- only COP EER using Eq. 4.9

. For an EER of 12.4, the ASHRAE 90.1

- 2010 level for this equipment class, this corresponds to a COP of 4.543. UEC is then calculated as

UEC 12

. 4 EER = 35 tons × [ 947 kW h / ton × 4

. 052 4

. 543 + 242 kW h / ton ] Eq. 4

. 11

A representative capacity for each equipment class was selected , as shown in Table 4.2

. 2 , based on the approximate average capacity of the units available in the market as identified in DOE’s market analysis

. Table 4.2

. 2 Water

- Cooled Air Conditioner Representative Capacities by Equipment Class Equipment Class Representative C ooling C apacity Water

- C ooled Air Conditioner, ≥65,000 and <135,000 Btu/h, Electric Resistance Heating or No Heating 8 t ons

Water

- C ooled Air Conditioner, ≥65,000 and <135,000 Btu/h, All Other Heating 8 t ons

Water

- C ooled Air Conditioner, ≥135,000 and <240,000 Btu/h, Electric Resistance Heating or No Heating 15 t ons

Water

- C ooled Air Conditioner, ≥135,000 and <240,000 Btu/h, All Other Heating 15 t ons

Water

- C ooled Air Conditioner, ≥240,000 Btu/h, Electric Resistance Heating or No Heating 35 t ons

Water

- C ooled Air Conditioner, ≥240,000 Btu/h, All Other Heating 35 t ons For this analysis, the annual supply fan power was not adjusted for different efficiency levels. A more detailed analysis of the engineering of higher efficiency water

- cooled air conditioners might show that some of the EER improvement might come from reducing supply fan power. One path to reducing fan power in rooftop equipment is believed to be thro ugh larger case designs and consequently lower internal fan pressure drop. However, based on the assumption that these water

- cooled air conditioner units are typically installed in mechanical rooms as opposed to rooftops, increased size may come at the exp ense of installation cost and available (rentable) floor space

. It is noted that for a given EER level, assuming that fan efficiency is improved to reach that level would mean less improvement to the compressor
condenser performance and less energy savings from the condens er portion of the packaged air conditioning system. The resulting UEC values for each class of water

- cooled air conditioner are provided in Table 4.2

. 3 for selected efficiency levels analyzed for the n ational e nergy s avings a nalysis

.4

- 11 Table 4.2

. 3 Water

- Cooled Air Conditioner Unit Energy Consum ption by Equipment Class Small Water

Cooled Air Conditioners Electric or No Heat 65

- 135 kBtu/h Small Water

Cooled Air Conditioners Other Heat 65

- 135 kBtu/h Large Water

Cooled Air Conditioners Electric or No Heat 135

- 240 kBtu/h Large Water

Cooled Air Conditioners Other Heat

135

- 240 kBtu/h Very Large Water

- Cooled Air Conditioners Electric or No Heat 240

- 760 kBtu/h Very Large Water

- Cooled Air Conditioners Other Heat

240

- 760 kBtu/h
Average Cooling Capacity ( tons ) 8 8 15 15 35 35 Efficiency Level ( EER ) Base Case

– Federal Standard 11.5 11.3 11.0 11.0 11.0 10.8 Efficiency Level 1 12.1 11.9 12.5 12.3 12.4 12.2 Efficiency Level 2 13.0 13.0 13.0 13.0 13.0 13.0 Efficiency Level 3 14.0 14.0 14.0 14.0 14.0 14.0 Efficiency Level 4 15.0 15.0 15.0 15.0 NA* NA* Efficiency Level 5

– “Max

- Tech”

– 16.4 16.4 16.1 16.1 14.8 14.8 Unit Energy Consumption ( kWh/yr ) Base Case

– Federal Standard 9,199 9,322 17,838 17,838 41,621 42,205 Efficiency Level 1 8,855 8,966 16,206 16,402 38,041 38,504 Efficiency Level 2 8,396 8,396 15,743 15,743 36,733 36,733 Efficiency Level 3 7,953 7,953 14,911 14,911 34,793 34,793 Efficiency Level 4 7,566 7,566 14,186 14,186 NA* NA* Efficiency Level 5

– “Max

- Tech”

– 7,101 7,101 13,490 13,490 33,422 33,422 *An E ff iciency L evel 4 was not identified for very large water

- cooled air conditioners.
4.2.2 Evaporatively Cooled Air Conditioners The analysis to assess the per unit energy use of evaporatively cooled air conditioners began with a review of the existing market (see c hapter 2 of the TSD). DOE did not identify any current models of evaporatively cooled air conditioners with less than 240,000 Btu/h cooling capacity. The review o f the market suggested that the evaporatively cooled equip ment models currently available are designed as packaged rooftop cooling units

. B ased on the available models identified , DOE estimated that the average capacity was 40 tons

. Therefore , DOE’s analysis of energy savings focused on typical applications for t his very large equipment class. Because of the large size, DOE believes that a common application would also be packaged VAV system serving a building

. To assess energy use for evaporatively cooled equipment, DOE modified the three

- story o ffice r eference b uilding spreadsheet model used for the water

- cooled equipment discussed previously to serve as the energy consumption model for the very large evaporatively cooled air conditioner equipment class. The EnergyPlus simulation tool can model evaporatively coo led unitary air conditioners directly, with only minor modifications from the air

- cooled unitary air conditioner equ ipment4- 12 models that were used in the original DOE medium office reference building model. EnergyPlus offers two alternatives for modeling evaporatively cooled air conditioners. One is to use performance curves that characterize the steady state efficiency as a function of ambient wet bulb temperature and indoor coil entering air wet bulb temperature (in contrast to the performance curves for air

- cooled equipment, which use ambient dry bulb temperature and indoor coil entering air wet bulb temperature). DOE was not able to derive separate performance curves for evaporatively cooled equipment as neither compressor nor overall compressor-condens er power data at different operating conditions were available in the manufacturer literature reviewed

. Therefore, DOE utilized the second approach provided for in EnergyPlus, which is to use performance curves for air

- cooled equipment that are a function of the condenser entering air dry bulb conditions, but provide for a evaporative condenser using an evaporative condenser effectiveness parameter that reduces the condenser coil entering air conditions to below the ambient air temperature according to the following equation :

T

c , i = T wb , o + (1

− Eff ) × ( T db , o − T wb , o ) E q. 4.12

Where :

T

c,i = temperature of the air entering the condenser coil (°F ), T

wb,o = wet bulb temperature of the outdoor air ( °F ), T

db,o = dry bulb temperature of the outdoor air ( °F ), and E

ff = evaporative condenser effectiveness

. DOE modified the medium office reference building model to use this latter modeling approach and defined an evaporatively cooled condenser for each of the t hree packaged VAV systems in the model

. An evaporative condenser with an effectiveness of 0.9, equivalent to the default value in EnergyPlus , was used

. DOE developed estimates for the compressor- condenser -only cooling COP based on the nominal rating condi tions . This was done by backing out the estimated fan power at nominal rating conditions and separately accounting for the impact of fan heat using Eq. 4.9 , as discussed for water -cooled equipment. The fan power ratio parameter R was estimated from a review of available manufacturer fan power data for an evaporatively cooled rooftop air conditio ner

. DOE estimated an R of 0.137 based on a review of available data for a large evaporatively cooled air conditioners and assuming a supply airflow rate of 350 cfm/ton of net cooling capacity. 7 For an 11.0 EER unit that DOE modeled directly, the calculate d compressor- condenser

- only COP was then estimated at 3.893. DOE used the fan power performance curves and peak fan power assumptions from the original reference medium office building model.

4.2.2.1 Baseline Annual Energy Use Per Ton – Evaporatively Cooled
DOE modeled energy use in the 15 climates identified previously using 11.0 EER evaporatively cooled air conditioners. Equipment was sized using design day sizing. DOE extracted the cooling capacity, annual equipment compressor-condenser energy consumption,4

- 13 and supply fan energy consumption for each evaporatively cooled air conditioner equipment unit simulated

. The annual energy consumption values were then normalized by dividing by the equipment capacity in cooling tons. The sum of the resulting c ompressor
- c ond enser energy per cooling ton and supply fan energy per cooling ton represents the annual energy consumption per cooling ton for equipment at that 11 EER efficiency level. These climate

- specific results, along with the climate weights , are shown in Table 4.2

. 4

. Table 4.2

. 4 Normalized Energy Use Ton by Climate for Evaporatively Cooled Air Conditioners, 11 .0 E nergy E fficiency R atio

Climate

Zone Representative Climate Location Construction Wt. Annual Compressor

Condenser

Energy kWh/ton Annual Fan Energy kWh/ton Annual System Energy kWh/ t on

1A Miami, F L 1.82% 1 , 683 380 2 , 068 2A Houston, TX 16.20% 1 , 293 329 1 , 626 2B Phoenix, AZ 4.99% 1 , 117 431 1 , 551 3A Memphis, TN 14.91% 979 305 1 , 287 3B El Paso, TX 10.15% 873 420 1 , 295 3C San Francisco, CA 1.82% 420 343 764 4A Baltimore, MD 18.18% 782 318 1 , 101 4B Albuquerque, NM 0.72% 620 405 1 , 027 4C Salem, OR 2.71% 426 332 759 5A Chicago, IL 16.96% 569 313 883 5B Boise, ID 5.70% 448 374 823 6A Burlington, VT 0.56% 473 340 814 6B Helena, MT 4.62% 335 356 692 7 Duluth, MN 0.56% 318 374 693 8 Fairbanks, AK 0.10% 186 383 569 Nat ional 100.00% 8 38 341 1 , 189
4.2.2.2 Energy Use Estimates for Baseline and Higher Efficiency Evaporatively Cooled Air Conditioners by Equipment Class Energy use consumptions were developed for representative evaporatively cooled equipment sizes corresponding to each equipment class , as was done for wate r
- cooled air conditioners. A representative capacity for each equipment class was selected , as shown in Table 4.2

. 5 , based on the approximate average c apacity of the units available in the market as identified in DOE’s market analysis.

Table 4.2

. 5 Evaporatively Cooled Air Conditioner Representative Capacities by Equipment Class Equipment Class Representative C ooling C apacity Evaporatively C ooled Air Conditioner, ≥ 240,000 Btu/h, Electric Resistance Heating or No Heating 40 t ons Evaporatively C ooled Air Conditioner, ≥ 240,000 Btu/h, All Other Heating 40 t ons

The national average energy consumption per ton figures from Table 4.2

. 4 were then multiplied by the selected equipment capacities for the evaporatively cooled equipment class4

- 14 an alyzed to establish the UEC values for the evaporatively cooled equipment classes at an 11 EER level. To assess the annual energy consumption at the other efficiency levels analyzed, DOE developed estimates of the compressor
- condenser

- only cooling COP for each other efficiency level using Eq. 4.9

. To establish annual energy consumption metrics for the higher efficiency levels, DOE used Eq. 4.10 , adjusting the baseline annual condenser energy consumption for each climate by the ratio of the baseline compressor
- condenser

- only cooling C OP at 11 EER to the compressor
- condenser

- only cooling COP at the efficiency levels analyzed

. The annual fan energy consumption estimates were held constant at the baseline level for all higher standards

. The UEC for each efficiency level analyzed is the su m of the annual compressor
- condenser energy consumption and the annual fan energy consumption

. The resulting UEC values for each class of evaporatively

- cooled air conditioner are provided in Table 4.2

. 6 for selected efficiency levels analyzed for the national energy savings analysi s.
Table 4.2

. 6 Evaporatively Cooled Air Conditioner Unit Energy Consumption Large Evaporatively

- Cooled Air Conditioner Electric or No Heat

240

- 760 kBtu/h Large Evaporatively

- Cooled Air Conditioner Other Heat 240

- 760 kBtu/h Average Cooling Capacity ( tons ) 40 40 Efficiency Level ( EER ) Base case 11.0 10.8 Level 1

– ASHRAE 11.9 11.7 Level 2 12.5 12.5 Max Tech 13.1 13.1 Unit Energy Consumption ( kWh/yr ) Base case 47,171 47,766 Level 1

– ASHRAE 44,732 45,243 Level 2 43,294 43,294 Max Tech 41,983 41,983
4.2.3 Computer Room Air Conditioners DOE estimated the per unit energy savings of computer room air conditioners using a temperature

- bin

- method analysis developed specifically for computer ro om air conditioners and implemented in a commercial spreadsheet. The spreadsheet calculates an annual UEC for air

- , water

- , or glycol

- cooled comput er room air conditioners of specific capacities based on the representative sizes analyzed.
The spreadsheet calculations presume that the heat load in the space being served is primarily sensible and that the energy consumption can be determined using the SCOP values for a given set of thermal load and outdoor conditions. The spreadsheet calculations further assume that conditions outside the server room environment have no effect on the thermal load ( i.e. , the thermal loads are a function of internal ga ins from the computer servers and associated4

- 15 equipment and do not take into account variations in envelope load or ventilation induced load that might occur through the year )

. The calculation proceed s in two principal stages. First , an hourly average powe r consumption performance map for a given class of computer room air conditioners ( at a defined efficiency level ) is developed using user entered design parameters for the equipment and either outdoor air dry bulb temperature or outdoor air wet bulb temper ature as the dependent variable , depending on equipment class. The performance map results are in terms of average hourly power usage and are calculated at the midpoint of defined 5 ° F temperature bins. Second, a dot

product multipl ica tion of the performan ce map (an average hourly power use versus temperature vector) and a vector of the annual hours associated with each temperature bin in a given climate is used to generate the expected annual energy consumption for that climate.

4.2.3.1 Development of Computer Roo m Air Conditioner Performance Map The calculation steps to represent the average hourly power consumption for a given temperature bin are based on the following analysis steps:

1. Estimate the average computer server power heat load served by a system

. 2. Estim ate the condensing fluid or dry bulb air temperature entering the CRAC condensing unit.

3. Estimate the load that would be met by an economizer (if incorporated ) for that hour

. 4. Calculate the thermal load served by the direct expansion ( DX ) system ( after accounting for any economizer load reduction)

. 5. Calculate the full load power consumption of the CRAC unit

. 6. Calculate the part

- load ratio for that hour (the ratio of the actual load to the full load capacity)

. 7. Calculate the part

- load factor for that hour (a n adjustment to the run time to reflect how the equipment responds at part

- load levels)

. 8. Calculate the hourly run fraction and the corresponding hourly average compressor
condenser power consumption for that hour

. 9. Add in the fan power , assuming continuous fan operation

. These steps are outlined in more detail below.

4.2.3.2 Estimate the A verage C omputer S erver H eat L oad CRAC units are designed for a given cooling capacity. However , due to sizing considerations and the fact that server loads will vary during the day or year, the ratio of the typical heat load in the space being served to the capacity of the CRAC unit will be less than 1.0.4

- 16 DOE e stimated the average sensible load on a CRAC unit to be 65 percent of the sensible capacity

. This value was based on an Australian Minimum Energy Performance Standard s ( MEPS ) study of close controlled air conditioners used for data processing applications , which used a 65 percent average load factor for this equipment ( operatin g on average at 65 percent of full load capacit y) based on industry consultation

. 8 DOE found that a different analysis presented to the California Energy C ommission (CEC) assumed that 25 percent of the load would be at each of the following load points (25, 50, 75, and 100 percent ), resulting in an ave rage load factor of 62.5 percent over the year , nearly equivalent to the 65 percent used in this analysis

. 9
There is likely significant variation in CRAC load factor s in the field. Server loads can vary d ue to computer usage, but more recently, c omputer o perating strategies like operating system virtu a lization can make it possible to more easily migrate computer server requirements between physical s ervers , allowing for individual computer servers to be placed in low activity states until needed. In addition, because of the requirements for high cooling reliability, purchase of redundant cooling units is common in the industry. Industry literature often speaks of purchasing n + 1 units for redundancy, where n is the number of units required to meet a design load and + 1 refers to the purchase of a redundant unit in case of system failure. One manufacturer refers to a general rule that the capacity of CRAC

- rated cooling must be 1.3 times the anticipated information technology load rating plus any cap acity added for redundancy , stating this approach works well with smaller network rooms of under 4,000 ft 2

. 10 However, DOE did not find any large studies that examined actual data on electrical load being cooled over the year relative to capacity installed, and used the 65 percent load factor figure for the energy analysis.

4.2.3.3 Estimate the C ondensing F luid or A ir T emperature for the Computer Room Air Conditioner U nit
The 15 classes of CRAC unit s can be grouped into three main categories based on condensing sys tem type : a ir

- cooled systems, water

- cooled systems, and glycol

- cooled systems. Air

- cooled systems utilize air

- cooled condensers. In most cases , the condensers are located outdoors in a split

- system arrangement where refrigerant piping connects the outdoor condenser to the indoor CRAC equipment

. I n some systems , outdoor air can be ducted into the building for condenser cooling. However , in both instances , the air used for cooling the condenser is at the prevailing outdoor air temperature

. Lower outdoor air temperatures during the year provide a cooler therma l sink and can improve the operating efficiency of the system by reducing the condensing temperature and corresponding refrigerant head pressure. However, for several reasons , including providing for oil migration and proper expansion valve performance, minimum head pressure requirements do exist in refrigerant system designs , including CRAC equipment designs. Various means of head pressure control exist in the air

- cooled condenser , including reducing fan speed, cycling of fans in multiple fan systems, and flooding compressors with liquid refrigerant to reduce the condenser volume that can be used for active refrigerant condensation.

Water

- and glycol

- cooled CRAC equ ipment utilize s either a water

- or a gly col

- cooled heat exchanger for condensing the refrigerant in the CRAC unit. These heat exchangers are constructed as an integral part of the CRAC unit. In the example of a water

- cooled unit, water4

- 17 enters the condenser and is heated by the condensing refrige rant. Warmer water th e n leaves the CRAC unit , is cooled , and flows back to the CRAC condenser. There are a variety of methods to cool the condensing water, including directly via dedicated open evaporative cooling towers serving one or more CRAC units, cen tral cooling towers serving both CRAC units as well as other heating, ventilating, and air

- conditioning ( HVAC ) equipment, closed evaporative cooling towers , and dry coolers. Glycol

- cooled CRAC unit s are designed similarly to water

- cooled CRAC unit s except that the refrigerant to fluid condenser is designed for a glycol

- water mix. The
w armed glycol

- water mix leaves the CRAC condenser and must be cooled prior to being returned to the CRAC unit. As with water cooling systems, a single glycol fluid loop can serve multiple CRAC units. However, this glycol

- cooling fluid loop is a closed system to prevent the loss of glycol. Thus , the fluid cooling is generally accomplished by a dry liquid

- to

- air heat exchanger, referred to in the industry as a dry cooler o r fluid cooler. One feature of glycol

- cooled systems is that the heat rejection system can readily operate at low ( below freezing ) temperatures since the freezing point of the glycol water mix is well below that of water

. Since CRAC units serve computer co oling loads that are year round, this is an important consideration.
As with air

- cooled CRAC units, a lower condenser fluid temperature can result in improved efficiency of the refrigeration system, down to the minimum head pressure allowed in the refrige ration system. Head pressure control in water

- and glycol

- cooled systems is typically accomplished by mixing or controlling the flow rate of the fluid in the condenser.
For the CR AC energy analysis tool, DOE made a number of generalized assumptions on how the fluid condenser temperature would be controlled:

• For air

- cooled systems, it was assumed that the condensing fluid (air) would be at the prevailing ambient dry bulb air temperature for a given hour.

• For water

- cooled systems, it was assumed that the en tering water would be provid ed by an open cooling tower with a constant 7 °F approach temperature ( i.e. , the temperature of the water returned to the CRAC unit would be 7 °F above the ambient wet bulb temperature )

. • For glycol

- cooled systems, it was assumed that the glycol

- water fluid would be cooled using a dry cooler and that the temperature of the fluid returned to the CRAC unit would be 10 °F above the ambient dry bulb air temperature. In addition, for each system , a “low t emperature” limit was also included , which helps establish the lower limit on the refrigerant head pressure. In water

- cooled systems, water must be held above the freezing temperature , and this sets a practical limit on the water temperature irrespective o f refrigerant head pressure considerations. In glycol

- and air

- cooled systems, th e

“ low temperature limit ” is defined not by the air or glycol properties , but by the need to maintain a minimum head pressure and thus condensing temperature of the refrigeran t. Since the heat exchangers are passive, active controls are used to limit either the heat transfer capabilities of the condenser ( e.g. , by reducing airflow or flooding condensers in air

- cooled equipment) or the condenser heat sink temperature ( through mi xing or flow rate control for glycol

- or water

- cooled equipment)

. The key use of these minimum temperatures for the spreadsheet model is to restrict the model such that that there is no improvement in CRAC efficiency with temperatures below these defined p oints. The control of the fluid temperature and the “minimum ’ s ” established in the model are shown in Table 4.2

. 7

.4

- 18 Table 4.2

. 7 Condensing C onditions E stablished for CRAC Equipment CRAC System Type Air

- Cooled Water

- Cooled Glycol

- Cooled Assumed Approach Temperature °F 0 7 10 Assumed Low Temperature Limit °F 40 40 35 In the model, weather data binned into 5 °F ambient dry bulb temperature bins are used to develop the annual performance for air

- cooled and glycol

- cooled CRAC equipment for different climates. W eather data in 5 °F wet bulb temperature bins is used to develop the annual performance for water

- cooled CRAC equipment

. 4.2.3.4 Economizer Load Reduction Economizers for DX cooling systems are system design features that allow for part or all of the thermal cooling load from the space to be removed prior to reaching the refrigerant evaporator coil. For CRAC equipment, the typical economizer is in corporated into the equipment
as a water

- to

- air or glycol

- to

- air cooling coil placed in the air stream prior to the DX evaporator coil. When weath er and system operations permit, the water or glycol economizer coil can be provided with cool entering fluid that can partially or wholly meet the cooling load from the space. This entering fluid temperat ure is the same as that provided to the CRAC unit from the external fluid cooling loop. R emaining heat load not met by the economizer coil can be met by the refrigeration coil. Because the economizer coil sits in the air stream , it results in an increase i n fan pressure and fan energy over the same equipment design without the economizer coil. The full load performance rating of equipment with economizer coils thus reflects this increased fan energy use.
T o estimate the load reduction for t he economizer coi l at a given hour, it was necessary to estimate both the entering fluid temperature and the overall heat transfer capability of the economizer coil. Limited information from manufacturer data indicates that , for CRAC systems with economizer coils, the econ omizers can generally meet the cooling load with fluid entering temperatures near 45 ° F. 11 , 12 , 13 In addition, ASHRAE 90.1

- 2010 14 has a requirement that evaporatively cooled water economizers serving primarily computer room s shall be able to meet 100 percent of the expected cooling load at 40 °F dry bulb / 35 °F wet bulb temperature outdoor conditions , and a system using dry

- cooler water economizers shall be able to meet 100 percent of the expected cooling load at 35 °F dry bulb outdoor conditions.

For CRAC classes using water or glycol economizers, DOE estimated the overall heat transfer coefficient (UA) for a fluid

- to

- air economizer cooling coil that could meet the net sensible cooling cap a city of the representative CRAC cooling unit at 75 °F return air te mperature and entering fluid conditions equal to the entering water or glycol conditions at 35 °F outdoor wet bulb temperature or 35 °F outdoor dry bulb temperature respectively , as follows:4

- 19 UA = Q s RAT

− EFT design Eq. 4

. 13

Where : Q

s = rated cooling capacity (sensible) , UA = o verall e conomizer c oil h eat t ransfer c oe f ficient (Btu/h

- F) , RAT = r eturn a ir t emperature (assumed equal to 75 °F) , and EFT design = e ntering f luid t emperature to e conomizer c oil (°F)

. For water

- cooled economizers, EFT design was estimated at 42 °F (35 °F outdoor wet bulb temperature + 7 °F approach temperature). For glycol

- cooled economizer, EFT design was estimated at 45 °F (35 °F outdoor dry bulb temperature + 10 °F approach temp erature). The economizer coil is assumed to be capable of operation any time the entering fluid is within a defined operating temperature range provided for by a low limit and high limit temperature control based on the entering water or glycol temperatur e. For water

- cooled economizers, the low limit set point was set to 35 ° F. For glycol

- cooled systems , the low limit set point of the range was set to

- 20 ° F, the lowest operating temperature provided for with a typical glycol loop. From a practical point of view, the economizer can meet 100 percent of the expected load under higher entering fluid temperatures than are defined by these conditions and will control the mixed fluid temperature in the coil to meet space conditions. A review of
m anufacturer lite rature suggested a range of economizer high limit controls set points, from systems where the economizer does not work when it can only partially meet the load to systems that would function with up to 65 °F entering fluid temperature. A temperature of 58 °F was chosen to represent a typical value within the range of high limit control strategies.
The heat load reduction from the economizer was calculated as the smaller of either the maximum potential heat transfer at the economizer coil for that hour or th e hourly computer server load : Q econ = Minimum [ UA × ( RAT

− EFT ) , Server Load ) ] Eq. 4

. 14

Where : Q econ = e conomizer load reduction , EFT = e ntering f luid t emperature to e conomizer c oil (°F) , and Server Load = s ensible heat load from the space being served. The server loads for the analysis were based on 65 percent of the sensible capacity , as discussed in section 4.2.3.2

. The analysis did not investigate the use of air side economizers. There appear to be questions within the industry on the use of air side economizers in spaces served by CRAC units. Air side economizers have significant pote ntial to reduce the annual load on all types of CRAC4

- 20 equipment as well as computer room air handler equipment using chilled water coils. Primary
concerns within the industry include the introduction of particulates, or chemical contaminants , in the supply air stream as well as the need to humidify the air entering through the economizer during cold weather when the outside air humidity ratio is low , which could lower the indoor space humidity conditions below a defined minimum humidity threshold . The cost o f any needed humidification depends on the type of humidifier ( e.g. , steam, ultrasonic), fuel costs, and the number of economizer hours and amount of humidification required for a given system and climate.

ASHRAE 90.1

- 2010 introduce d tables of economizer requirements for cooling systems in server computer rooms as a function of climate and system cooling capacit y (greater than 65,000 Btu/h) , but at the same time provided significant exceptions that limited the application of the econo mizer requirements

. These exceptions include ( 1) any building where the total design cooling load of all computer rooms in the building is less than 3,000,000 Btu/h and the building in which they are located is not served by a centralized chilled water pla nt ; ( 2) individual computer rooms where the total design cooling load is less than 600,000 Btu/h and the building in which they are located is served by a centralized chilled water ; ( 3) spaces where less than 600,000 Btu/h of computer room cooling capacity is being added to an existing building ; and ( 4) computer space s that can be defined as serving specific critical needs. DOE had no data on which to establish the frequency of air side economizer usage with the different classes of CRAC equipment , but note s that the exemptions provided in ASHRAE 90.1

- 2010 would seem to indicate that their use would be limited to very large computer server applications

. DOE notes that t he analysis assumption of no air side economizer usage likely sets an upper bound on the
c ooling loads and average energy use of air

- cooled CRAC equipment.

4.2.3.5 Estimate the Cooling Load The cooling load at any hour is estimated as the computer server load plus the heat from operating CRAC supply fans, minus the load met by the economizer coil in that hour. Base fan power was estimated at 0.418 W /cfm for systems without an economizer coil. Fan airflow was estimated at 472 cfm per ton of sensible capacity based on review of manufacturer data for a variety of CRAC products. For systems with an ec onomizer coil, the fan power was increased by the amount necessary to increase the total system power and reduce the EER of a corresp onding baseline water

- or glycol

- cooled system by the 0.05 EER reduction seen in the ASHRAE 90.1 requirements. The estimate d fan power f or water

- cooled equipment with economizers was 0.461, 0.464, and 0.467 W/cfm for less than 65,000 Btu/h, greater than or equal to 65,000 Btu/h but less than 240,000 Btu/h, and greater than or equal to 240,000 Btu/h CRAC equipment , respectively

. The estimated fan power for glycol

- cooled equipment with economizers was 0.46 4 , 0. 477 , and 0. 480 W/cfm for less than 65,000 Btu/h, greater than or equal to 65,000 but less than 240,000 Btu/h, and greater than or equal to 240,000 Btu/h CRAC equipment , res pectively. It is assumed that all of the fan power is converted to heat that must be removed by the CRAC system.

4.2.3.6 Estimate the Hourly Power Usage To estimate the average energy consumed for any hour of operation , it was assumed that the supply fan would run in any hour, but that the average hourly compressor power could be4

- 21 reduced from the nominal rating condition due to reduced compressor run time (corresponding to reduced cooling load) or external conditions. Compressor power and run time were estimated based on methods outlined in the EnergyPlus building simulation user manuals f or DX cooling equipment . DOE developed three curves to modify the full load performance (express ed in terms of the e nergy i nput r atio (EIR)

— the ratio of compressor

- condenser pow er to the cooling output at the evaporator coil, in consistent units) as a function of condenser air or entering fluid temperatures. DOE had no detailed equipment performance data from manufacturers on which to base these relationships. DOE instead used ex ample tables presented in ASHRAE Standard 127 informative appendix D , 15 which show the variation of full load SCOP as a function of outside air temperature (for air

cooled products) or as a function of entering fluid temperature (for glycol

- cooled products)

. DOE converted the SCOP da ta to sensible EIR data (ratio of compressor

- condenser power to sensible capacity) using the following formula :

EIR s = 1

− R SCOP + R Eq. 4

. 15

Where :

EIR s = s ensible e nergy i nput r atio for compressor
- c ondenser (kW/kW) , and R = r atio of fan power to total system power

. The fan power ratio used for these initial calculation s was approximately 0.2. Four SCOP points, corresponding to four o utside air temperatures , were converted t o sensible EIR points for air

- cooled equipment

. d These sensible EIR values were then normalized by dividing by the EIR at the nominal 95 °F rating condition, and the normalized values plotted against outdoor air temperature. A second order polynomial curve fit using the outdoor temperature as the independent variable was used to characterize the relationship. This resulting value for the curve fit is referred to as COOL

- EIR

- FT (cooling EIR as a function of temperature). A similar analysis of the ASHRAE 127 a ppendix D EIR v ersus entering fluid temperature for glycol

- cooled equipment was also conducted. The curve fits are provided in Table 4.2

. 9 and shown in Figure 4.2

. 2 for air

- and glycol

- cooled equipment.

ASHRAE 127 a ppendix D d oes not provide similar data for water

- cooled equipment as a function of condenser entering fluid tempera ture. For water

- cooled equipment, DOE developed a COOL

- EIR

- FT curve fit parallel to that of the glycol

- cooled systems, but which passed through the ASHRAE 127 nominal rating point for water

- cooled CRAC equipment. This curve fit is also shown in Figure 4.2

. 2

.

d O nly points corresponding to var iable capacity equipment were provided in ASHRAE 127 a ppendix D for these temperatures , but the ratings were full load ratings points. Non

- variable capacity equipment are not required to test performance at other than the nominal outside air temperature or entering fluid conditions.4

- 22 Figure 4.2

. 2 Cooling EIR as a Function of Entering Condenser Fluid Air Temperature

Table 4.2

. 8 CRAC Performance as a F unction of E ntering Condenser Fluid or O utdoo r Air T emperature Equipment Energy Input Ratio Performance Curve Air

- Cooled COOL

- EIR

- FT =

- 0.000006 x OAT 2 + 0.004941 x OAT + 0.584620 Water

- Cooled COOL

- EIR

- FT = 0.00001566 x EFT 2 + 0.00065099 x EFT + 0.79467383 Glycol

- Cooled COOL

- EIR

- FT = 0.0000228 x EFT 2

– 0.0001787 x EFT +0.7700841 OAT = o utside a ir t emperature (oF) , EFT = e ntering f luid t emperature (oF) These relationships were used to estimate the full load sensible EIR for a given outdoor air or entering condenser fluid temperature. The actual power consumption for a given hour reflects the full load EIR at the operating conditions as well as the cooling load at that hour. DOE defines the part

- load ratio (PLR) for a given hour as the ratio of the actual load to cooling capacity at each hour. DOE calculated the PLR for each temperature bin. DOE also calculated a part

- load factor (PLF) correlation curve for each CRAC unit. The product of the rated EIR and Cool

- EIR

- FT modifier curve is divided by the calculated PLF to give the “effective” EIR for a given simulation hourly bin

. In EnergyPlus, t he PLF correlation is designed to account for efficiency losses, primarily due to compressor cycling. EnergyPlu s provides a typical PLF correlation curve for single compressor DX system as

PLF = 0

. 85 + 0

. 15 × PLR Eq. 4

. 16 COOL

- EIR

- FT_air cooled =

- 0.000006 OAT 2 + 0.004941 OAT + 0.584620 R2 = 0.997371

COOL

- EIR

- FT_glycol = 0.0000228 EFT 2

- 0.0001787 EFT + 0.7700841 R2 = 0.9973387 COOL

- EIR

- FT_water cooled = 0.00001566 EFT 2 + 0.00065099 EFT + 0.79467383 R2 = 0.99992938

0.00 0.20 0.40 0.60 0.80 1.00 1.20

0 20 40 60 80 100 120EIR (T) / EIR (Nominal)

Entering Condenser Fluid or Air Temperature (F) Air Cooled Unit Glycol Cooled Unit
Water Cooled Unit Poly. (Air Cooled Unit) Poly. (Glycol Cooled Unit) Poly. (Water Cooled Unit)4

- 23 Where :

PLF = p art

- l oad f actor (single compressor systems) , and PLR = part

- load ratio as defined in Eq. 4.4

. DOE developed a second average PLF relationship for two

- compressor systems by assuming that two equally sized compressors were use d. F or a PLR of less than 0.5 , one compressor is off and one compressor cycles to meet the load. For a PLR greater than 0.5 , one compressor operates continuously without cycling and the second compressor cycles to meet the additional load. DOE developed se parate PLF relationships for each compressor by examining the load on each compressor relative to the capacity of each compressor ( assuming each provided 50 percent of the total capacity of the system). The overall PLF for two

- compressor systems is the loa d

- weighted average of the PLF for each single compressor in the system , and is defined by

PLR ≤ 0

. 5 , PLF = 0

. 85 + 0

. 15 × 2 × PLR PLR > 0

. 5 , PLF = 1 PLR × ( 1 2 + ( PLR

− 0

. 5 ) × ( 0

. 85 + 0

. 15 × 2 × ( PLR

− 0

. 5 ) ) ) Eq. 4

. 17
The run time fraction (RTF) for the compressor

- condenser system in a given hour bin is defined as

RTF = PLR PLF Eq. 4

. 18 The average power for the compressor

- condenser in a given bin is calculated as the product of the cooling capacity, the adjusted EIRs, and the run time fraction for that bin, or

Power cond = CAP s × EIR s × COOL

- EIR

- FT × RTF Eq. 4

. 19

Where :
CAP s = the nominal sensible capacity of the CRAC unit

. The total system power is the sum of the fan power and the compressor

- condenser power,
or

Power = Power cond + Power supply fan s Eq. 4

. 20

Where :

Power supply fan = the supply fan power , assuming continuous operation

.4

- 24 Performance maps of estimated CRAC hourly power consumption for each temperature bin are provided in Appendix 4 B by equipment class

. 4.2.3.7 Calculate Annual Energy Consumption by Cli mate and by State The annual energy consumption for each representative equipment selection in a given climate is calculated as the dot product of the performance map vector for that CRAC unit as a function of outdoor temperature and the corresponding bin hour vector for a given climate. DOE developed dry bulb and wet bulb temperature bin hours for 2 39 U.S. climates using Typical Mete o rological Year 3 ( TMY3 ) weather data , and from these developed 2 39 location

- specific estimates of annual energy consumption f or each class of equip ment at each efficiency level. These location

- specific energy use estimates were converted to state

- level energy use estimates by equipment class at each efficiency level using previously developed TMY

- location weighting factors. See A ppendix 4C for a discussion of the development of these climate weights. e The resulting estimates of annual energy consumption by state are shown in Appendix 4B and referenced direct ly in the life

- cycle cost spreadsheet discussed in chapter 6

. 4.2.4 Water

- Sour ce Variable Refrigerant Flow Heat Pumps
Annual UEC estimates for VRF water source heat pumps at or greater than 135,000 Btu/h cooling capacity were developed based on whole building energy simulation of a medium sized prototype office building in 1 5 locati ons around the United States , corresponding to each of 15 ASHRAE climate zones discussed previously. 3 The prototype office building model is a two

story, 20,000 ft 2 building that has five distinct HVAC zones, a core zone , and four perimeter zones, per floor. The building model used ASHRAE 90.1

- 2004 compliant construction to represent relatively recent construction practices and provide loads suitable for representing both new and replacement/retrofit construction. Assumptions as to the envelope design and building form factor were drawn from previous analysis o f small office buildings. 16 The HVAC model was modified to use two separate water

- source VRF systems, one serving each floor, in conjunction with a gas fired boiler and a singl e cooling tower to serve the condensing water loop for the VRF systems. The simulation tool was a commercial version of the DOE2.1E building simulation tool, with the capability to model water source VRF equipment via custom DOE2.1E functions

. 17 This simula tion tool utilizes actual performance curves from equipment manufacturers for a large number of VRF units, including water

- source VRF units. Results from simulations of water

- source VRF heat pumps a t three different efficiency levels in each of the 1 5 loca tions were developed and aggregated to estimates of national energy use for VRF systems using commercial construction weights developed for each climate. These national results in turn were used to estimate the annual energy consumption of the VRF units at five different efficiency levels that DOE analyze d

. Details of this analysis follow.

4.2.4.1 Building Model Characteristics The small office building prototype model uses characteristics typical of buildings of this size and use. The building is a 20,000 ft 2
, two

- story, square building. The square shape was

e Original location weights developed for TMY2 locations. TMY3 weather data for these locations was used in the analysis of computer room air conditioner energy use.4

- 25 originally chose n to provide an orientation

- neutral simulation platform

. The building has a window

- area ratio of 21 percent with a 1 2

- ft floor

- to

- floor height (a plenum is not directly modeled) . The small of fice building is assumed to have slab

- on
- grade construction for the floor, mass walls, and a flat roof. A picture of the building structure is shown in Figure 4.2

. 3

. The building is divided into five HVAC zones, four perimeter zones , and a core zone, per floor

.

Figure 4.2

. 3 Small Office Building Design View Where applicable, the building components for this model were assumed to “just meet” the minimum prescriptive requirements of ASHRAE Standard 90.1

- 2004. Envelope and window performance characteristics are shown in Table 4.2

. 9 for each ASHRAE climate zone

. Table 4.2

. 9 Envelope Performance Characteristics Climate

Zone Characteristic Climate Wall U

- Factor Roof U

- Factor Glass U

- F actor Window SHGC / SC
1A Miami , FL 0.58 0.063 1.2 0.25 / 0.287 2A Phoenix , AZ 0.58 0.063 1.2 0.25 / 0.287 2B Houston , TX 0.58 0.063 1.2 0.25 / 0.287 3A Memphis , TN 0.58 0.063 0.57 0.25 / 0.287 3B El Paso , TX 0.58 0.063 0.57 0.25 / 0.287 3C San Francisco , CA 0.151 0.063 1.2 0.39 / 0.448 4A Baltimore , MD 0.151 0.063 0.57 0.39 / 0.448 4B Albuquerque , NM 0.151 0.063 0.57 0.39 / 0.448 4C Salem , OR 0.151 0.063 0.57 0.39 / 0.448 5A Chicago , IL 0.123 0.063 0.57 0.39 / 0.448 5B Boise , ID 0.123 0.063 0.57 0.39 / 0.448 6A Helena , MT 0.104 0.063 0.57 0.39 / 0.448 6B Burlington , VT 0.104 0.063 0.57 0.39 / 0.448 7 Duluth , MN 0.09 0.063 0.57 0.49 / 0.563 8 Fairbanks , AK 0.08 0.048 0.46 0.49 / 0.563 * *No SHGC requirements for ASHRAE C l imate Z one 8 ; however , selected values represent simulation inputs and are equivalent to Z one 7. The building is assumed to follow typical small office occupancy patterns, with peak occupancy occurring from 8 a.m. to 5 p.m. weekdays and limited occupancy from 6 a.m. un til 8 a.m. and from 5 p.m. until midnight for janitorial functions. The building is assumed to be4

- 26 unoccupied on weekends and holidays. Schedules for lighting and miscellaneous equipment were default schedules for office buildings selected using the EnergyP ro user interface for ASHRAE 90.1 compliance using a simulation performance approach option . These schedules reflect a 5

- day per week occupancy , with principal occupancy between 8 a.m. and 6 p.m. and limited additional occupancy in the evenings for janitor ial services. HVAC schedules followed occupancy with the HVAC system scheduled on from 6 a.m. to 9 p.m. at night. Thermostat set points were 70 ° F heating, 75 °F cooling during occupied hours , and 55 ° F heating, 99 ° F cooling on weekends and holidays

. Ther e is a 2

- hour thermostat warm

- up period between 6 a.m

. and 8
a .m

. weekday mornings. Figure 4.2

. 4 through Figure 4.2

. 7 show the schedules for lighting, plug loads, and occupancy

. Figure 4.2

. 7 provides the building space t emperature set points. The first three of these schedules represent multipliers to defined daily “peak” conditions. In the case of ventilation, it is assumed that ventilation occurs whenever the fan system is operational. For the occupancy, lighting, plug use, and ventilation, t he peak values were defines as follows . Lighting density was defined using the ASHRAE 90.1

- 2004 maximum lighting power density requirement of 1.0 W/ft 2 . Plug load density ( e.g. , computers and other miscellaneous loads) was 1.0 W/ft 2 . Occupancy peak density was defined as 88 persons for the building ( 4.4 persons per 1,000 ft 2 ). Total outdoor air exchange during the occupied period (effective ventilation) was based on ASHRAE Standard 62

- 1999 and was set to 0.1 3 cfm/ft 2 of floor area fo r each conditioned zone

.

Figure 4.2

. 4 Small Office Lighting Schedule 0% 10% 20% 30% 40% 50% 60% 70% 80% 90% 100%

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24Percent of Peak Power Density

Weekday Sat/Sun/Hol4

- 27 Figure 4.2

. 5 Small Office Plug Load Schedule

Figure 4.2

. 6 Small Office Occupancy Schedule 0% 10% 20% 30% 40% 50% 60% 70% 80% 90% 100%

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24Percent of Peak Power Density

Weekday Sat/Sun/Hol

0% 10% 20% 30% 40% 50% 60% 70% 80% 90% 100%

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24Percent of Peak Occupancy

Weekday Sat/Sun/Hol4

- 28 Figure 4.2

. 7 Small Office Thermostat Set Point Schedule

4.2.4.2 Variable Refrigerant Flow Modeling Characteristics DOE developed estimates of the key equipment input parameters , including fan power and compressor
- condenser COP at rating conditions , to represent systems at three different efficiency levels identified in the AHRI certified product directory for VRF multi

- split air conditioners and heat pu mps. 18 These efficiencies corresponded to ( 1) the lowest efficiency level identified in the directory and close to the ASHRAE baseline ; ( 2) an efficiency level corresponding to the highest efficiency identified for ducted systems ; and ( 3) an efficiency level near the highest efficiency identified for ductless systems. The AHRI 1230

- 2010 19 test procedure used for these ratings provides that each condensing unit be tested both as a ducted system, representing equipment using indoor units that are connected to short distribution ducts, and as a ductless system, representing equipment using ductless indoor sections that provide conditioned air directly to the building space served. Because of a higher external static pressure when testing ducted units, the rat ed efficiency (EER and COP) of a given condensing unit is lower when tested as a ducted system than when tested as a ductless system. The two higher efficiency levels simulated utilized the same condensing unit but represent ratings as a ducted and as a du ctless system , respectively. As the ratings data do not identify the indoor units used, DOE selected a representative ducted indoor section and developed supply fan power estimates based on that unit for ducted systems representing the first two efficienc y levels simulated. For non

- ducted systems where there was a large variety of indoor sections available, DOE developed an average fan power estimate based on average supply fan power data for five different ductless indoor section designs. Key characterist ics of the VRF units simulated are shown in Table 4.2

. 10

. 0 10 20 30 40 50 60 70 80 90 100

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24Thermostat Setpoint (°F)

Heat-Weekday Heat-Sat/Sun/Hol Cool-Weekday Cool-Sat/Sun/Hol4

- 29 Table 4.2

. 10 VRF Modeling Characteristics
System

Type Model *,** Nominal Ratings Condens ing Unit Performance † Indoor Section Classification Indoor Section Fan Power W/cfm EER COP COPc ool COPh eat Without Heat Recovery A 10.3 3.95 4.546 5.039 Ducted 0.2408 B 11.8 4.51 5.366 5.942 Ducted 0.2408 C 14.4 4.47 5.366 5.942 Non

- d ucted 0.1136 With Heat Recovery A 10.3 3.95 4.546 5.039 Ducted 0.2408 B 11.8 4.51 5.366 5.942 Ducted 0.2408 C 14.4 4.47 5.366 5.942 Non

- d ucted 0.1136 * ASHRAE 90.1

- 2010 specifies a minimum efficiency of 10.0 EER for water source VRF heat pumps > 135,000 Btu/h without heat recovery and a 9.8 EER for water source VRF heat pumps > 135,000 Btu/h with heat recovery. ** Model A data was for the lowest efficiency model shown in the AHRI database for this equipment class. Model B data was fo r the model with the highest efficiency “ducted” rating from the same manufacturer (a second manufacturer had data at 11.9 EER for ducted equipment). Model C data represents the Model B unit tested as a ductless unit and thus has the same compressor perfor mance. A second manufacturer had a 14.5 EER unit , which represented the maximum shown in the database, but with a lower heating COP. † Based on EnergyPro simulation tool data and used as entry in DOE2.1E building description language (BDL) data files creat ed by EnergyPro. COPcool referrers to compressor
- condenser cooling COP at rated conditioner as utilized by EnergyPro . COPheat refers to compressor
- condenser heating COP at rated condition. Compressor performance data checked against manufacturer literatu re showing compressor output capacity and power consumption. DOE performed simulations of the prototype office building using these three VRF efficiency levels and using weather data for 1 5 specific cities. E ach specific city location provided a representative climate corresponding to one of the ASHRAE 90.1

- 2004 climate
zones

. A corresponding location

- specific TMY3 data file was used as the source for hourly weather data

. The location s simulated are shown in Table 4.2

. 11

. Table 4.2

. 11 Climate and Locations for Simulations ASHRAE 90.1 Climate Zone City Construction Wt. 1A Miami, FL 3.24% 2A Houston, T X 15.22% 2B Phoenix, A Z 2.98% 3A Atlanta, G A 15.03% 3B Los Angeles, C A 10.08% 3C San Francisco, C A 1.61% 4A Baltimore, M D 19.29% 4B Albuquerque, NM 0.52% 4C Seattle, WA 2.98% 5A Chicago, IL 19.37% 5B Denver, C O 4.34% 6A Minneapolis, MN 4.21% 6B Helena, M T 0.57% 7 Duluth, M N 0.51% 8 Fairbanks, A K 0.06% DOE first performed preliminary design and annual simulation runs to size the indoor equipment and two outdoor units serving the conditioned zone s

. Because each VRF manufacturer has limits on the overall ratio of indoor unit capacity to outdoor unit capacity , the4

- 30 sum of the capacity of the indoor units was not allowed to be greater than 150 percent of the outdoor unit system capacity for any VRF system.
Six sets of s imulation were done , with each set utilizing all 15 climates: three sets for the

“ without heat recovery ” equipment , corresponding to the three identified efficiency levels shown in Table 4.2

. 10 , and three sets for equipment with heat recovery at the same efficiency levels. Th e annual electrical energy use for the VRF equipment, including each condensing unit and all associated evaporator units, was extracted from the si mulation results for each building simulated and normalized by cooling capacity to provide annualized estimates of total VRF cooling, heating, and fan energy consumption at the average cooling capacity estimated by DOE for VRF system s

. For VRF systems with out heat recovery , DOE estimated the average capacity at 216,000 Btu/h based on the average for available equipment found in the AHRI certified products directory. For VRF systems with heat recovery, DOE estimated the average capacity at 192,000 Btu/h usin g the same data source. DOE estimated the national average energy use for VRF systems at both the ASHRAE 90.1

- 2010 efficiency level and the max

- tech level identified for ducted VRF systems on a national basis using commercial building construction weights previously developed by DO E and assigned to each of the 15 climates. Because VRF can be used in many types of buildings and building designs, t he construction weights used were those for all commercial buildings

. 6 These weights are shown in Table 4.2

. 11

. Table 4.2

. 12 shows the calculated natio nal average energy use from the simulations for heating, cooling, and fan energy
on a per
- ton of capacity installed basis

. Annual fan energy was later apportioned between heating and cooling for accounting purposes using the ratio of the heat ing to cooling annual energy consumpt ion

. Table 4.2

. 12 National Average VRF Energy Use

– Simulations Results
Type Simulation Level Cooling kWh/ton Heating kWh/ton Fan kWh/ton Fan

Cooling kWh/ton Fan

Heating kWh/ton
No Heat Recovery A 827 331 365 261 104 B 729 295 365 260 105 C 729 295 172 122 49
Heat Recovery A 786 211 365 288 77 B 683 180 365 289 76 C 683 180 172 136 36 For each equipment class, DOE developed a piecewise linear relationship between the national average cooling energy use and the reciprocal of the cooling EER for the three efficiencies modeled. DOE also developed a piecewise linear relationship between the national average heating energy use and the reciprocal of the heating COP for the three efficiencies modeled. DOE then used these relationships to estimate the annual average cooling and heating energy use at the ASHRAE baseline efficiency level f and at f our higher efficiency levels, including the highest EER and COP levels found in the AHRI certified product directory for each equipment class (identified as max

- tech levels for this analysis). These levels are shown in Table 4.2

. 13

. Level 2 corresponded to the highest efficiency found for ducted VRF equipment.

f No water

- source VRF equipment greater than 135 kBtu/h with efficiency levels lower than the ASHRAE 90.1

- 2010 efficiency levels were identified in the AHRI certified product directory or through manufacturer literature. Consequently, the lowest efficiency levels examined in the energy use characterization were the ASHRAE 90.1

- 2010 efficiency levels.4

- 31 DOE held the fan energy use co nstant for levels at and below L evel 2 to that estimated based on the ducted VRF simulations. Level 1 was approximately midway between the ASHRAE baseline and Level 2. The highest efficiency level was L evel 4 , which is considered the max
- tech level and was based on the highest efficiency found in the AHRI certified directory. Level 4 c orresponded to a ductless system

. DOE estimated the energy use at the max

- tech level using the linear relationship between the higher two efficiencies simulated

. Level 3 was intermediate between L evel s 2 and 4. Annual energy use at L evel 3 was calculated b ased on interpolation between L evel 2 and the max

- tech level. In all, DOE developed annual energy consumption estimates for efficiency levels at EER values of 10.0, 11.0, 12.0, 13.0, and 14.5 for water

- source VRF heat pumps without heat recovery. DOE devel oped annual UEC estimates for efficiency levels at EER values of 9.8, 11.0, 12.0, 13.0, and 14.5 for water

- source VRF heat pumps with heat recovery. While ASHRAE 90.1

- 2010 provides for a 0.2 EER difference between the EER for products with and without heat recovery, a review of all manufacturer ratings for these two equipment classes in the AHRI certified directory did not show evidence of any difference in rated EER for product models with and without heat recovery (identified by models having the same or nearly the same base model number)

. The resulting annual UEC for each equipment class at each efficiency level analyzed is shown in Table 4.2

. 13

. Table 4.2

. 13 Unit Energy Consumption Estimates by Efficiency Level for Water

- Source VRF Equipment Equipment Class es > 135 kBtu/h Water

Source VRF Without Heat Recovery >135 kBtu/h Water

Source VRF with Heat Recovery Representative Capacity 216 kBtu/h 192 kBtu/h
Efficiency Level EER/COP UEC per Standard Level kWh/yr Cooling Heating Cooling Heating

ASHRAE Standard 10.0/4.0

– VRF without heat recovery 19,998 7,786 NA NA
9.8/3.9

– VRF with heat recovery NA NA 17,830 4,634
Level 1 11.0 / 4.1 18,693 7,607 16,353 4,450 Level 2 12.0 /4.3 17,606 7,440 15,348 4,309 Level 3 13.0 /4.4 16,496 7,026 14,265 4,030

Max Tech 14.5/4.6

– VRF without heat recovery 15,031 6,414 NA NA
14.5/4.7

– VRF with heat recovery NA NA 12,836 3,621
COP = coefficient of performance; EER = energy efficiency ratio; kBtu = thousand British thermal units; UEC = unit energy consumption; VRF = variable refrigerant volume ; NA = Not Applicable

.4

- 32 REFERENCES

1

. Carrier Corporation. OMNIZONETM Indoor Self

- Contained Systems . Farmington, CT.
< http://xpedio.carrier.com/idc/groups /public/documents/marketing/04

- 850275

- 25.pdf >
2

. Briggs, R.L., R.G. Lucas, and Z.T. Taylor. Climate Classification for Building Energy Codes and Standards: Part 1

– Development Process

. ASHRAE Transactions . 2003. 1: pp. 4610.

3

. Briggs , R.L ., R.G. Lucas, and Z.T. Taylor. Climate Classification for Building Energy Codes and Standards: Part 2

– Zone Definitions, Maps, and Comparisons

. ASHRAE Transactions . 2003. 1: pp. 4611

4 U.S. Department of Energy

− Office of Energy Efficiency and Renewable Energy. EnergyPlus Energy Simulation Software , V 6 .0. 20 11 . Washington, D.C.
< http://apps1.eere.energy.gov/buildings/energyplus >
5

. Carrier Corpora tion. Carrier Product Data Remote Air

- Cooled and Water

- Cooled Indoor Self

- contained Systems and Water Source Heat Pumps . 50BV

- 4PD.pdf. (Last accessed March 2011.)
< http://www.commercial.carrier.com/commercial/hvac/carrier/0,3068,CLI1_DIV12_ETI12 690_MID4412,00.html >
6

. Jarnagin , R.E. and G.K. Bandyopadhyay

. Weighting Factors for the Commercial Building Prototypes Used in the Development of ANSI/ASHRAE/IESNA Standard 90.1

- 2010

. 2010. Pacific Northwest National Laboratory: Richland, WA. Report No. PNNL

19116. < http://www.pnl.gov/main/publications/e xternal/technical_reports/PNNL

19116.pdf >
7

. Trane. Packaged IntelliPakTM Rooftop Systems 20 to 130 tons . Product catalog for Packaged Rooftop Air Conditioners IntellipakTM

– S*HL,S*HK. (Last accessed April 2011.) < http://www.trane.com/COMMERCIAL/DNA/View.aspx?i=1102 >
8

. EnergyConsult Pty Ltd. Equipment Energy Efficiency Committee Regulatory Impact Statement Consultation Draft: Minimum Energy Performa nce Standards and Alternative Strategies for Close Control Air Conditioners . September 2008. Jindivick, Australia. Report No

. 2008/11. < http://www.energyrating.gov.au >
9

. California Utilities State Wide Codes and Standards Team. 2013 California Building Energy Efficiency Standards Codes and Standards Enhancement Initiative Draft Measure Implementation Template

– Data Centers

. March 2011.
< http://www.energy.ca.gov/title24/2013standards/prerulemaking/documents/2011

- 04

- 11_workshop/review/Data_Center_33011.pdf >
10

. Rasmussen, N

. Calculating Total Cooling Requirements for Data Cen ters

. American Power Conversion Corporation, West Kingston, RI. White Paper 25.
< http://www.apcmedia.com/salestools/NRAN

- 5TE6HE_R3_EN.pdf >4

- 33 11

. Liebert Corporation. Engineering White Paper: Utilizing Economizers Effectively in the Data Center . 2003. Columbus, O H.

12

. Emerson Network Power . Economizer Fundamentals: Smart Approaches to Energy

Efficiency Free

- Cooling for Data Centers

. 2010. Columbus, OH.

13

. Data Aire, Inc. Data Aire Se ries Air Cooled, Water/Glycol Cooled 6 through 30 ton Dual Circuits

. Orange , C A

. < http://www.dataaire.com/wp

content/uploads/2011/11/DataAireSeries

- DX

- R410A.pdf >
14

. American Society of Heating, Refrigerating and Air

- Conditioning Engineers. ASHRAE 90.1

- 2010, Energy Standard for Buildings Except Low

- Rise Residential Buildings . 2010. Atlanta, GA.

15

. American Society of Heating, Refrigerating and Air

- Conditioning Engineers

. ASHRAE 127 , Method of Testing for Rating Computer and Data Processing Room Unitary Air Conditioners . 2007. Atlanta, GA.

16

. Thornton, B.A., W. Wang, Y. Huang, M.D. Lane, B. Liu. Technical Support Document: 50% Energy Savings for Small Office Bui ldings

. 2010. Pacific Northwest National Laboratory: Richland, WA. Report No. PNNL

- 19341.

17

. EnergySoft LLC. Users Manual: EnergyPro Version 5

. 2011. Novato, CA.

18

. Air

- Conditioning, Heating, and Refrigeration Institute. Certified Product Directory: Vari able Refrigerant Multi

- Split Air

- Conditioners and Heat Pumps: AHRI Standard 1230

. 2011 Edition. Effective date: September 16, 2011. Arlington, VA.
< http://www.ah ridirectory.org/ahridirectory/pages/vrf/VRFDirectory_20110916.pdf >
19

. Air

- Conditioning, Heating, and Refrigeration Institute

. AHRI 1230

- 2010 , Performance Rating of Variable Refrigerant Flow (VRF) Multi

- Split Air

- Conditioning and Heat Pump Equipment . 2010. Arlington, VA.