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 Conditioner 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 Characteristics............................... 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 ................................... 4-18
4-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 ........................................................ 4-28
4-1

CHAPTER 4. ENERGY USE CHARACTERIZATION

4.1 INTRODUCTION

The national energy savings characterization described in this chapter provides estimates

of the energy savings consumers would realize from the establishment of standards 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 more 

energy efficient standards for the same equipment classes.

This chapter describes the energy use analysis for water-cooled and evaporatively cooled 

commercial packaged air conditioners, computer room air conditioning (CRAC) equipment, and 

water-source variable refrigerant flow (VRF) equipment with a cooling capacity greater than 

135,000 British thermal units per hour (Btu/h). These equipment types are described in separate 

subsections of this chapter, which detail the analysis and the determination of baseline annual per

unit energy consumption (UEC) estimates for specific equipment classes within each of the three

equipment types examined. This chapter 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: energy efficiency ratio (EER) for cooling efficiency, 

coefficient of performance (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 

chapter 8 of the TSD. The review suggested that most water-cooled air conditioner units on the 

market are designed for installation inside commercial buildings (as opposed to on building 

rooftops). The shipments data suggested that, in recent years, shipments were dominated by 

larger equipment (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. Manufacturer 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 data developed from computer simulations of a prototypical, medium 

sized, commercial office building. The prototype building model was a three-story, 53,600 ft2
4-2

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

a

In this building model, referred to as a “medium office,” each floor is served by a separate 

packaged air conditioning unit. The hourly data used for this analysis was previously developed 

from simulations using this model and the DOE EnergyPlus building simulation software, and 

reflected building simulations in 15 climate locations in the United States.

b Each 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 subsequently 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 Appendix 4A.

The medium office reference building model used 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 from 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, these 

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 equipment 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 an 

existing 25-ton water-cooled air conditioner.

5 Cooling capacity and combined compressorcondenserc 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 modeling of 

the water-cooled condenser and supply fan power follows.

 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 (PNNL) 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 efficiency) 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 Performance

EnergyPlus provides two methods of modeling equipment performance of water-source 

heat pumps. The method used for this work is referred to as the “equation-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. Least squares 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 volumetric 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 variations from the implementation in 

EnergyPlus included the use of inch-pound 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:

Qtotal

Qtotal_ref

= A1 + A2 × � Twb

Twb_ref� + A3 × � Tw,inTw,in_ref� + A4 × � V̇

air

V̇

air_ref� + A5 × � V̇

w

V̇

w_ref�

Eq. 4.1

Qsen

Qsens_ref

= B1 + B2 × � Tdb

Tdb_ref� + B3 × � TwbTwb_ref� + B4 × � Tw,inTw,in_ref� + B5 × � V̇

air

V̇

air_ref� + B6 × � V̇

w

V̇

w_ref�

Eq. 4.2

Power

Powerref

= C1 + C2 × � Twb

Twb_ref

� + C3 × � Tw,in

Tw,in_ref� + C4 × � V̇

air

V̇

air_ref� + C5 × � V̇

w

V̇

w_ref�

Eq. 4.3

Where:

A1 through A5 = equation fit coefficients for the total cooling capacity,

B1 through B6 = equation fit coefficients for the sensible cooling capacity,

C1 through C5 = equation fit coefficients for the power consumption,

Qtotal = cooling capacity (Btu/h),

Qtotal_ref = cooling capacity at reference condition (Btu/h),

Qsens = sensible cooling capacity (Btu/h),

Qsens_ref = sensible cooling capacity at reference condition (Btu/h),

Power = compressor-condenser power (W),

Powerref = compressor-condenser power at reference condition (W),

Twb = entering air wet bulb temperature (Rankine (°R)),

Twb_ref = entering air wet bulb temperature at reference temperature (°R),

Tdb = entering air dry bulb temperature (°R),

Tdb_ref = entering air dry bulb temperature at reference temperature (°R),
4-4

Tw,in = entering water temperature (°R),

Tw,in_ref = entering water temperature at reference temperature (°R),

V̇

air = volumetric airflow rate (cubic feet per minute (cfm)),

V̇

air_ref = volumetric airflow rate at reference (cfm),

V̇

w = volumetric water flow rate (gallons per minute (gpm)), and

V̇

w_ref = volumetric 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 performance 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 they 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 = C1 + C2 × PLR

Eq. 4.4

Where:

Cool-PLF-FPLR = cooling power part-load factor modifier,

PLR = ratio of cooling load to cooling capacity under the conditions of interest, and

C1 and C2 = equation fit coefficients for the part-load performance curves.

C1 and C2 values of 0.771 and 0.229, respectively, were used based on the default partload 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:

PowerHourly = Power ×

PLR

Cool-PLF-FPLR

Eq. 4.5

Where:

PowerHourly = hourly average power in response to hourly cooling load,

Power = steady state power at conditions as defined in Eq. 4.3,

PLR = part-load ratio as defined in Eq. 4.4, and

Cool-PLF-FPLR = part-load modifier as defined in Eq. 4.4.

For 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-water 
4-5

flow rates were available, and for the modeling purpose the 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 Temperature

The performance equations developed in this spreadsheet 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 temperature 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

Tw = minimum�Twb,o + 7 , 70� + 459.67

Eq. 4.6

Where:

Tw = building supply water temperature from cooling tower (°R), and

Twb,o = outdoor 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 condition 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 conditioner 

units for this building model.

4.2.1.3 Supply 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:

Powerfan,max = ∆P × Q

6350 × nf × nm × 0.746

Eq. 4.7

Where:

Powerfan,max = design fan power (kW),

ΔP = total fan static pressure (in. H20),

Q = design supply airflow rate (cfm),

nf = fan efficiency, and

nm = motor and drive efficiency.
4-6

A design total fan static pressure of 4.35 in. H2O, fan efficiency of 0.65, and motor 

efficiency 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. H2O versus 5.35 in. H2O) 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 inside 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 

adjusted 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:

PLFfan = C1 + C2 × FF + C3 × FF2 + C4 × FF3+C5 × FF4

Eq. 4.8

Where: 

PLFfan = a part-load factor multiplier equal to the ratio of system fan power to design fan power,

FF = fraction of system design airflow rate, and

C1 through C5 = equation coefficients for the VAV fan power performance curve.

The coefficients C1 through C5 used for the VAV fan power performance curve are those 

from the original packaged VAV system used in the reference building model and are as follows:

C1 = 0.040759894

C2 = 0.08804497

C3 = -0.07292612

C4 = 0.943739823

C5 = 0

4.2.1.4 Water-Cooled Compressor-Condenser-Only Cooling Coefficient of 

Performance vs. Energy Efficiency Ratio Relationship

For analysis of energy use at the base efficiency (EER) level, DOE developed estimates 

for the compressor-condenser-only efficiency (compressor-condenser-only cooling COP) for the 

water-cooled air conditioner based on the nominal rating conditions for an 11 EER unit.

Compressor-condenser-only cooling COP at rating conditions is an input to both the EnergyPlus 

simulation and DOE’s spreadsheet models and is the ratio of the gross output divided by the 

compressor-condenser power, in consistent units. This parameter can be calculated by backing 

out the estimated fan power at nominal rating conditions from the rated energy consumption 

while accounting for the effect of fan heat. This is done by calculating the compressorcondenser-only cooling COP according the following equation: 
4-7

COPCU_Cool =

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

COPCU_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 rated 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 calculated 

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 fan energy use for each of the three watercooled air conditioner units in the reference 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 the office buildings represented in this analysis compared with 

that of all 15 DOE reference buildings types with construction floor space tallied in the study.
4-8

Figure 4.2.1 Comparison of Floor-Space Weighting Factors by Region, Small and Medium 

Office Buildings versus All Buildings

In addition, Table 4.2.1 shows a parameter called the “full load equivalent operating 

hours” (FLEOH), which is calculated as the annual cooling load served, divided by the system 

cooling capacity. In this case, the FLEOH shown is the result from weighting of the FLEOH 

calculated for each individual air conditioning unit by the relative capacity of that unit to the total 

capacity of all units serving the reference building model. While not used in DOE’s energy 

calculation, 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 8

Relative Fraction of Construction by Climate Region

DOE Climate Region

All Buildings

Office Buildings
4-9

Table 4.2.1 Normalized Energy Use per Ton by Climate for Water-Cooled Air 

Conditioners, 11.0 Energy Efficiency Ratio

Climate 

Zone

Representative 

Climate Location

Construction 

Wt.

Annual 

CompressorCondenser 

Energy 

kWh/ton

Annual Fan 

Energy 

kWh/ton

Annual 

System 

Energy 

kWh/ton

FLEOH

hours

1A Miami, FL 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

National 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 Equipment by Equipment Class.

National average estimates of the energy use for efficiencies other than those directly 

modeled in the spreadsheet tool were developed by first scaling the national average compressorcondenser 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,

UECEER = CAP × [CondUECn ×

COPCU_Cool_Modeled

COPCU_Cool_EER

+ FanUECn]

Eq. 4.10

Where:

CAP = nominal equipment capacity in cooling tons,

UECEER = unit energy consumption at the specific EER level being evaluated,

CondUECn = normalized condenser energy consumption from the modeling, kWh/ton – national

basis, 

COPCU_Cool_Modeled = the cooling compressor-condenser-only COP used in the spreadsheet 

equipment model,
4-10

COPCU_Cool_EER = the cooling compressor-condenser-only COP corresponding to the specific 

EER level being evaluated, and

FanUECn = normalized compressor-condenser energy consumption from the modeling, kWh/ton 

– national.

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 

COPEER 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

UEC12.4 EER = 35 tons × [947 kWh/ton ×

4.052

4.543 + 242 kWh/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 Cooling Capacity

Water-Cooled Air Conditioner, ≥65,000 and <135,000 Btu/h, Electric 

Resistance Heating or No Heating

8 tons

Water-Cooled Air Conditioner, ≥65,000 and <135,000 Btu/h, All Other 

Heating

8 tons

Water-Cooled Air Conditioner, ≥135,000 and <240,000 Btu/h, Electric 

Resistance Heating or No Heating

15 tons

Water-Cooled Air Conditioner, ≥135,000 and <240,000 Btu/h, All Other 

Heating

15 tons

Water-Cooled Air Conditioner, ≥240,000 Btu/h, Electric Resistance 

Heating or No Heating

35 tons

Water-Cooled Air Conditioner, ≥240,000 Btu/h, All Other Heating 35 tons

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 through 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 expense 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 compressorcondenser performance and less energy savings from the condenser 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 national energy savings analysis.
4-11

Table 4.2.3 Water-Cooled Air Conditioner Unit Energy Consumption by Equipment Class

Small WaterCooled Air 

Conditioners

Electric or No 

Heat

65-135 kBtu/h

Small WaterCooled Air 

Conditioners

Other Heat

65-135 kBtu/h

Large WaterCooled Air 

Conditioners

Electric or No 

Heat

135-240 

kBtu/h

Large WaterCooled 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 Efficiency Level 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 chapter 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 of the market suggested that the evaporatively cooled equipment models 

currently available are designed as packaged rooftop cooling units. Based 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 this 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 office reference building 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 cooled unitary air conditioners 

directly, with only minor modifications from the air-cooled unitary air conditioner equipment 
4-12

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-condenser 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: 

Tc,i =Twb,o + (1 − Eff) × (Tdb,o − Twb,o)

Eq. 4.12

Where: 

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

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

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

Eff = 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 three 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 conditions. 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 conditioner. 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 calculated 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 compressor-condenser 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 Energy Efficiency Ratio

Climate 

Zone

Representative Climate 

Location

Construction 

Wt.

Annual 

CompressorCondenser 

Energy 

kWh/ton

Annual 

Fan 

Energy 

kWh/ton

Annual System 

Energy 

kWh/ton

1A Miami, FL 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

National 100.00% 838 341 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 water-cooled air 

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

4.2.5, based on the approximate average capacity 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 Cooling Capacity

Evaporatively Cooled Air Conditioner, ≥ 240,000 Btu/h, Electric 

Resistance Heating or No Heating 40 tons

Evaporatively Cooled Air Conditioner, ≥ 240,000 Btu/h, All Other 

Heating 40 tons

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 class 
4-14

analyzed 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 COP 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 sum 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 

analysis.

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 room air conditioners and 

implemented in a commercial spreadsheet. The spreadsheet calculates an annual UEC for air-, 

water-, or glycol-cooled computer 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 gains from the computer servers and associated 
4-15

equipment and do not take into account variations in envelope load or ventilation induced load 

that might occur through the year).

The calculation proceeds in two principal stages. First, an hourly average power 

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 temperature 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 dotproduct multiplication of the performance 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 Room 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. Estimate 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 (an 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 compressorcondenser 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 Average Computer Server Heat Load

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 estimated 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 Standards

(MEPS) study of close controlled air conditioners used for data processing applications, which 

used a 65 percent average load factor for this equipment (operating on average at 65 percent of 

full load capacity) based on industry consultation.

8 DOE found that a different analysis presented 

to the California Energy Commission (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 average 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 factors in the field. Server loads can 

vary due to computer usage, but more recently, computer operating strategies like operating 

system virtualization can make it possible to more easily migrate computer server requirements 

between physical servers, 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 capacity added for redundancy, 

stating this approach works well with smaller network rooms of under 4,000 ft2.

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 Condensing Fluid or Air Temperature for the Computer 

Room Air Conditioner Unit

The 15 classes of CRAC units can be grouped into three main categories based on 

condensing system type: air-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. In 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 thermal 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 equipment utilizes either a water- or a glycol-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, water 
4-17

enters the condenser and is heated by the condensing refrigerant. Warmer water then 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, central 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 units are designed similarly to water-cooled 

CRAC units except that the refrigerant to fluid condenser is designed for a glycol-water mix. The 

warmed 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 or 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 cooling 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 refrigeration 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 CRAC 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 entering water would be provided

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 temperature” 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 of refrigerant head pressure considerations. In glycol- and air-cooled systems, the

“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 refrigerant. 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 mixing 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 points. 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 Conditions Established 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. Weather 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 incorporated 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 weather 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 temperature is the same as that provided to the CRAC unit from the 

external fluid cooling loop. Remaining 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

in 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.

To estimate the load reduction for the economizer coil 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 economizers can generally meet the cooling load with fluid entering 

temperatures near 45 °F.11,12,13 In addition, ASHRAE 90.1-201014 has a requirement that 

evaporatively cooled water economizers serving primarily computer rooms 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 capacity of the representative CRAC cooling unit at 75 °F return air temperature 

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 = Qs

RAT − EFTdesign

Eq. 4.13

Where:

Qs = rated cooling capacity (sensible),

UA = overall economizer coil heat transfer coefficient (Btu/h-F),

RAT = return air temperature (assumed equal to 75 °F), and

EFTdesign = entering fluid temperature to economizer coil (°F).

For water-cooled economizers, EFTdesign was estimated at 42 °F (35 °F outdoor wet bulb 

temperature + 7 °F approach temperature). For glycol-cooled economizer, EFTdesign was 

estimated at 45 °F (35 °F outdoor dry bulb temperature + 10 °F approach temperature).

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 temperature. 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 

manufacturer literature 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 the hourly computer 

server load:

Qecon = Minimum[UA × (RAT − EFT), Server Load)]

Eq. 4.14

Where:

Qecon = economizer load reduction,

EFT = entering fluid temperature to economizer coil (°F), and

Server Load = sensible 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 potential to reduce the annual load on all types of CRAC 
4-20

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 of 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 introduced tables of economizer requirements for cooling systems

in server computer rooms as a function of climate and system cooling capacity (greater than 

65,000 Btu/h), but at the same time provided significant exceptions that limited the application of 

the economizer 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 plant; (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 spaces 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 notes 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 the analysis assumption of no air side economizer usage likely sets an upper bound on the 

cooling 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 economizer coil, the fan power was increased by 

the amount necessary to increase the total system power and reduce the EER of a corresponding 

baseline water- or glycol-cooled system by the 0.05 EER reduction seen in the ASHRAE 90.1 

requirements. The estimated fan power for 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 0.464, 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, respectively. 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 be 
4-21

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 for DX cooling equipment. DOE developed three 

curves to modify the full load performance (expressed in terms of the energy input ratio (EIR)—

the ratio of compressor-condenser power 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 example 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 aircooled products) or as a function of entering fluid temperature (for glycol-cooled products). DOE 

converted the SCOP data to sensible EIR data (ratio of compressor-condenser power to sensible 

capacity) using the following formula:

EIRs = 1 − R

SCOP + R

Eq. 4.15

Where:

EIRs = sensible energy input ratio for compressor-condenser (kW/kW), and

R = ratio of fan power to total system power.

The fan power ratio used for these initial calculations was approximately 0.2. Four SCOP 

points, corresponding to four outside air temperatures, were converted to 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 appendix D EIR versus 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 appendix D does not provide similar data for water-cooled equipment as a 

function of condenser entering fluid temperature. 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 Only points corresponding to variable capacity equipment were provided in ASHRAE 127 appendix 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 Function of Entering Condenser Fluid or Outdoor Air 

Temperature

Equipment Energy Input Ratio Performance Curve

Air-Cooled COOL-EIR-FT = -0.000006 x OAT2 + 0.004941 x OAT + 0.584620

Water-Cooled COOL-EIR-FT = 0.00001566 x EFT2 + 0.00065099 x EFT + 0.79467383

Glycol-Cooled COOL-EIR-FT = 0.0000228 x EFT2 – 0.0001787 x EFT +0.7700841 

OAT = outside air temperature (oF), EFT = entering fluid temperature (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, the PLF correlation is designed to account for efficiency losses, primarily 

due to compressor cycling. EnergyPlus 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 OAT2 + 0.004941 OAT + 0.584620

R2 = 0.997371

COOL-EIR-FT_glycol = 0.0000228 EFT2 - 0.0001787 EFT + 0.7700841

R2 = 0.9973387

COOL-EIR-FT_water cooled = 0.00001566 EFT2 + 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 80 100 120

EIR (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 = part-load factor (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 used. For 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 separate 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 load-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

Powercond = CAPs × EIRs × COOL-EIR-FT × RTF

Eq. 4.19

Where:

CAPs = 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 = Powercond + Powersupply fans

Eq. 4.20

Where:

Powersupply 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 4B by equipment class.

4.2.3.7 Calculate Annual Energy Consumption by Climate 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 239 U.S. climates using Typical 

Meteorological Year 3 (TMY3) weather data, and from these developed 239 location-specific 

estimates of annual energy consumption for each class of equipment 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 Appendix 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 directly in the life-cycle cost spreadsheet discussed in chapter 6.

4.2.4 Water-Source 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 15 locations around the United States, corresponding to each of 15 

ASHRAE climate zones discussed previously.3 The prototype office building model is a twostory, 20,000 ft2 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 of 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 single 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 simulation 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 at three different efficiency levels in each of the 15

locations 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 analyzed. 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 ft2

, 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 chosen to provide an orientation-neutral simulation platform. The building has a 

window-area ratio of 21 percent with a 12-ft floor-to-floor height (a plenum is not directly 

modeled). The small office 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-Factor 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 Climate Zone 8; however, selected values represent simulation inputs and are equivalent to Zone 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. until 8 

a.m. and from 5 p.m. until midnight for janitorial functions. The building is assumed to be 
4-26

unoccupied on weekends and holidays. Schedules for lighting and miscellaneous equipment were 

default schedules for office buildings selected using the EnergyPro 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 janitorial 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. There 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 temperature 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, the 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/ft2

. Plug load density (e.g., computers and other miscellaneous 

loads) was 1.0 W/ft2

. Occupancy peak density was defined as 88 persons for the building (4.4

persons per 1,000 ft2

). Total outdoor air exchange during the occupied period (effective 

ventilation) was based on ASHRAE Standard 62-1999 and was set to 0.13 cfm/ft2 of floor area 

for 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 24

Percent of Peak Power Density

Weekday

Sat/Sun/Hol
4-27

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 24

Percent 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 24

Percent of Peak Occupancy

Weekday

Sat/Sun/Hol
4-28

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 pumps.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 1230-201019 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 rated 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 ductless 

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 efficiency 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 characteristics 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 24

Thermostat Setpoint (°F)

Heat-Weekday

Heat-Sat/Sun/Hol

Cool-Weekday

Cool-Sat/Sun/Hol
4-29

Table 4.2.10 VRF Modeling Characteristics

System 

Type Model*,**

Nominal 

Ratings

Condensing Unit

Performance† Indoor 

Section 

Classification

Indoor 

Section Fan 

Power 

W/cfm EER COP COPcool COPheat

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-ducted 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-ducted 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 for 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 performance. 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 created 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 literature 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 15 specific cities. Each 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 locations 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, TX 15.22%

2B Phoenix, AZ 2.98%

3A Atlanta, GA 15.03%

3B Los Angeles, CA 10.08%

3C San Francisco, CA 1.61%

4A Baltimore, MD 19.29%

4B Albuquerque, NM 0.52%

4C Seattle, WA 2.98%

5A Chicago, IL 19.37%

5B Denver, CO 4.34%

6A Minneapolis, MN 4.21%

6B Helena, MT 0.57%

7 Duluth, MN 0.51%

8 Fairbanks, AK 0.06%

DOE first performed preliminary design and annual simulation runs to size the indoor 

equipment and two outdoor units serving the conditioned zones. Because each VRF

manufacturer has limits on the overall ratio of indoor unit capacity to outdoor unit capacity, the 
4-30

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 simulation 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.

The annual electrical energy use for the VRF equipment, including each condensing unit and all 

associated evaporator units, was extracted from the simulation 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 systems. For VRF systems without 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 using 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 DOE and assigned to each of the 15 climates. Because VRF can be used 

in many types of buildings and building designs, the 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 national 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 heating to cooling annual energy

consumption.

Table 4.2.12 National Average VRF Energy Use – Simulations Results

Type Simulation 

Level

Cooling 

kWh/ton

Heating 

kWh/ton

Fan 

kWh/ton

FanCooling 

kWh/ton

FanHeating 

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 levelf and at four 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 constant for levels at and below Level 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 Level 4, which is considered 

the max-tech level and was based on the highest efficiency found in the AHRI certified directory.

Level 4 corresponded 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 Levels 2 and 4. Annual energy use at Level 3 was calculated based on 

interpolation between Level 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 developed 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 Classes

>135 kBtu/h WaterSource VRF Without 

Heat Recovery

>135 kBtu/h WaterSource 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, V6.0. 2011. Washington, D.C. 

<http://apps1.eere.energy.gov/buildings/energyplus>

5. Carrier Corporation. 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. PNNL19116. <http://www.pnl.gov/main/publications/external/technical_reports/PNNL19116.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 Performance 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 Centers. 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, OH.

12. Emerson Network Power. Economizer Fundamentals: Smart Approaches to EnergyEfficiency Free-Cooling for Data Centers. 2010. Columbus, OH. 

13. Data Aire, Inc. Data Aire Series Air Cooled, Water/Glycol Cooled 6 through 30 ton Dual 

Circuits. Orange, CA.<http://www.dataaire.com/wpcontent/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 Buildings. 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: 

Variable Refrigerant Multi-Split Air-Conditioners and Heat Pumps: AHRI Standard 1230.

2011 Edition. Effective date: September 16, 2011. Arlington, VA. 

<http://www.ahridirectory.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.