Performance Evaluation of Rooftop Air Conditioning Units
At High Ambient Temperatures
Ramin Faramarzi, Bruce Coburn, Rafik Sarhadian, Scott Mitchell, and R. Anthony Pierce,
Southern California Edison
ABSTRACT
The cooling capacity of packaged, air-cooled, rooftop air conditioning units (RTUs)
declines as the ambient temperatures increases. Rooftop air conditioning units’ performance
parameters such as power consumption and efficiency rating are much less understood at
temperatures exceeding 115oF. These temperatures are typically beyond those at which the RTU
manufacturers test their products. Many climate regions within California achieve these high
ambient temperatures on a consistent basis.
The purpose of this research is to determine the impact of high ambient temperatures on
the electric demand and cooling efficiency of five-ton RTUs. This study seeks to better
understand efficiency degradations and electric demand implications of high efficiency and
standard efficiency units for three leading RTU manufacturers under realistic peak summer
cooling conditions seen in California.
A series of laboratory tests were performed to quantify the impacts of high ambient
temperatures on the energy efficiency and power usage of three standard and three high
efficiency, five-ton RTUs. Performance of these units was evaluated and compared at 85oF,
95oF, 105oF, 115oF, 120oF, 125oF and 130oF air temperatures measured at the condenser air inlet.
Test results revealed that in some instances the compressor of the high efficiency model
experienced a more drastic increase in electric demand than the standard efficiency model,
requiring up to 74% more power per degree Fahrenheit of change in ambient temperature. At an
ambient temperature of 130oF, the power demand of one of the high efficiency units surpassed
that of the standard model by 120 watts. The high efficiency unit from a second manufacturer
would have consumed more power than its standard model had it not failed to operate beyond an
ambient temperature of 127oF. Despite these instances of a high efficiency unit having a higher
electric demand, all three high efficiency units proved more efficient than the standard models.
Overall, the high efficiency units maintained superior efficiency and cooling output than the
standard efficiency units.
Each RTU’s cooling capacity deteriorated as the ambient temperature increased. The
standard efficiency units used more power than the high efficiency units under almost all
conditions. The rate of cooling capacity deterioration, however, varied between individual units.
Generally, the high efficiency units provided superior cooling capacity to that of their standard
efficiency counterparts. The superior capacity was due to improvements such as larger
evaporator and condenser surface areas. Coupling the higher capacities with efficient
compressors in high efficiency units also contributed to higher energy efficiency ratios (EER).
Introduction
Packaged air conditioning units are used to cool 36.5 billion square feet of commercial
space in the United States with an average annual energy consumption of 16.0 kWh per square
3-52
foot, which puts a heavy burden on the electric grid (EIA 1999). A large percentage of Heating,
Ventilating, and Air-Conditioning (HVAC) equipment used in California’s small-commercial
sector consists of packaged, rooftop air conditioners of five-ton capacity (Johnson 2001).
The Seasonal Energy Efficiency Ratio (SEER) is a common cooling efficiency
performance indicator of these units. The SEER provides the ratio of the total cooling of an
RTU in Btus during its normal usage period for cooling to the total electric energy input in watthours, (ARI 210/240). While SEER provides an indication of the cooling efficiency, it fails to
address the power or electric demand implications of different RTUs. The closest index
commonly used in the industry that implies the electric demand efficiency of the air conditioning
unit is the Energy Efficiency Ratio (EER). Manufacturers determine the EER of their equipment
by dividing the net cooling capacity in Btu/hr by the total electric input measured in watts with
95oF ambient air at the condenser inlet. Noteworthy, the 95oF ambient temperature is much
milder than the actual climate and site conditions that exist in the inland valley and desert regions
of California. In this study, units with SEER of 12 and higher are referred to as “high efficiency”
and those with SEERs of 10 as “standard” models.
The dependency of the EER on cooling capacity variations often obscures the variation in
power requirements when either the load on the evaporator or the inlet condenser air temperature
changes. At increasing ambient temperatures, the EER of the unit may remain reasonably steady
because the rate at which its cooling capacity degrades is much less than the rate of increase in
power consumption. The principle questions and issues that led to the investigations described
in this paper were:
1.
2.
3.
4.
Do higher SEER RTUs require more power (kW) than standard SEER models at ambient
temperatures greater than 115°F?
Does the cooling capacity of a high SEER unit degrade faster than a standard SEER
model at high ambient temperatures?
Absence of reliable and independent third party performance test data at high ambient
temperatures.
Frequent occurrence of temperatures above Air-Conditioning and Refrigeration
Institute’s (ARI) 115oF test condition in California combined with the fact that condenser
coil inlet air temperatures are typically higher than ambient dry-bulb temperatures.
Approach
Standard and high efficiency RTUs were purchased from three leading HVAC
manufacturers for testing in this project. A series of laboratory tests were performed at Southern
California Edison’s (SCE’s) Refrigeration and Thermal Test Center (RTTC) to quantify the
impacts of high ambient temperatures on the performance and power use of standard and high
efficiency five-ton RTUs. Performance of these units was evaluated and compared at 85oF,
95oF, 105oF, 115oF, 120oF, 125oF and 130oF air temperatures at the condenser air inlet.
A comprehensive data acquisition system was developed to monitor 138 channels of data
every 20 seconds and log critical RTU performance parameters. The entering air conditions to
the evaporator, which was one of the critical control points, was maintained at 80oF dry-bulb and
67oF wet-bulb for all test scenarios, regardless of changes in the ambient temperature. The
collected data allows for a quantitative analysis of the operation of each RTU. Tests were
carried out following a strict test protocol developed by the RTTC staff and based on ARI
3-53
210/240 (ARI 1989). Furthermore, an uncertainty analysis model was developed to establish
confidence in obtained results. Preliminary results show that uncertainty in the refrigerant-side
cooling capacity is less than 3%. Power measurements and EER calculations contain
uncertainties of 0.5% and 4%, respectively. These uncertainty values are comparable to those
found in similar studies (LeRoy 1997).
Description of Tested Units
Specifications of all six RTUs utilized for testing at the ambient conditions mentioned
previously are shown in Table 1. All three high efficiency units came equipped with scroll
compressors rated at higher efficiencies and in most cases contain larger coil surface areas.
None of the high efficiency units was equipped with variable speed compressors.
Table 1. Manufacturer Specifications of Standard and High Efficiency 5-Ton RTUs
Manufacturer:
Type:
A
B
Standard
High
Efficiency
5
60,500
(17.7)
57,500
(16.9)
6.50
C
Standard
High
Efficiency
Standard
High Efficiency
5
63,000
(18.5)
60,000
(17.6)
5.50
5
63,100
(18.4)
60,000
(17.6)
6.78
5
62,400
(18.3)
59,500
(17.4)
5.56
5
60,375
(17.7)
57,000
(16.7)
6.70
5
64,375
(18.9)
61,000
(17.9)
5.55
10.0 / 8.8
13.0 / 11.0
10.2 / 8.9
12.0 / 10.7
10.0 / 8.0
13.0 / 11.0
7.875 lbs.
10 lbs.
4.9 lbs.
8.4 lbs.
6.875 lbs.
10 lbs.
Reciprocating
Scroll
Scroll
Scroll
Reciprocating
Scroll
15.8
17.7
13.6
14.2
12.2
15.8
Cooling Performance
Nominal Tonnage
Gross Cooling Capacity - Btu/hr
(kW)
Net Cooling Capacity - Btu/hr
(kW)
Total Unit Power (kW)
SEER (Btu/W-hr)/ EER
(Btu/hr/W)
Refrigerant Charge Furnished
(HCFC-22)
Compressor
Type
Rated EER @ 95°F
(based on compressor mfg data)
Condenser Coil
Net Face Area (sq. ft.)
14.6
14.6
8.81
10.96
10.42
16.5
Number of Rows / Fins per Inch
2 / 20
2 / 20
2 / 17
3 / 17
2 / 17
2 / 17
Condenser Fan
Motor Horsepower (W)
1/3 (248)
1/3 (248)
1/3 (248)
1/3 (248)
.25 (190)
.25 (190)
Motor speed (RPM)
1075
1075
1075
1075
1100
1100
Total Motor Power (W)
360
360
360
360
325
320
Total Air Volume (cfm)
Evaporator Coil
Net Face Area ( sq. ft.)
4,200
4,200
3,470
3,370
4,000
4,100
6.25
6.25
5.00
7.71
5.50
5.50
Number of Rows / Fins per Inch
2 / 15
3 / 15
3 / 16
4 / 16
Balanced Port
TXV
Balanced Port
TXV
Standard Short
Orifice TXV
Short Orifice
TXV
3 / 15
Special
Evaporatorintegrated
Orifice Device
4 / 15
Special
Evaporatorintegrated Orifice
Device
1.5 hp
(1.1 kW) 208/230V
1.5 hp
(1.1 kW) 208/230V
1 hp
(0.61 kW) 208/230V
1 hp
(0.61 kW) 208/230V
1.30 hp
(.989 kW) 208/230V
1.20 hp
(.895 kW) 208/230V
11 ½ x 9
11-1/2 x 9
11 x 11
11 x 11
10 x 10
10 x 10
Expansion Device Type
Evaporator Blower and Drive
Selection
Nominal Motor Output (Power –
voltage)
Wheel Nominal Diameter x
Width (in.)
3-54
Test Design
Test Protocols
The determination of all test units’ capacities and performance characteristics closely
followed test methods specified in ARI 210/240-89, which adopts the American Society of
Heating, Refrigeration, and Air-conditioning Engineers (ASHRAE 1988) Standard 37-1988.
Cooling capacities were determined by applying the air-enthalpy and the refrigerant-enthalpy
methods to test results. All following performance comparison discussions rely on results
obtained from the refrigerant-enthalpy method.
All tests were performed at constant conditions for a period of one hour. Each rooftop
unit to be tested was installed in the test room in accordance with the manufacturer’s installation
instructions. Ambient air velocity in the vicinity of the condenser section was monitored closely
and maintained below 500 fpm. The ten foot high ceiling in the outdoor control environment
room provided more than six feet of clearance for condenser discharge air flow. A distance of at
least three feet was provided between the test room’s walls and the RTU case.
ARI 210/240 stipulates that the cooling capacity rating includes the effects of blower fan
heat, but excludes supplementary heaters. Power measurements include input to the compressor
and fans, control power, and any other items required as part of the RTU’s normal operation.
Ambient temperatures of 95°F and 115°F, two of the standard ARI test conditions, were
performed in this project to benchmark test results against the manufacturer’s published data.
ARI 210/240 requires that, when multiple analyses are used, the total cooling capacity
shall be the evaporator-side capacity of two simultaneously conducted analyses, which shall
agree within 6.0%. Gross cooling capacity was determined by the product of refrigerant mass
flow rate and change in refrigerant enthalpy across the evaporator coil. The sensible capacity
was obtained by subtracting the latent load, which was based on mass of condensate, from the
gross cooling capacity.
Airflow over the evaporator coil was maintained at 1,875 (± 5%) cfm against a total static
pressure of 0.55 (± 0.10) in.wg. for all test conditions. Each unit was charged with the weight of
refrigerant prescribed by the manufacturer plus an additional amount to compensate for the extra
volume of the mass flow meter and associated piping.
Data Acquisition
The RTTC test facility is equipped with a National Instruments SCXI computer-based
data acquisition system to acquire and log test data. It includes a high-performance signal
conditioning and instrumentation system for PC-based data acquisition and control. The data
acquisition system processes and averages 100 records from 138 data points every 20 seconds.
This factory calibrated system is traceable to the National Institute of Standards and
Technology’s (NIST) standards. Prior to each test, all instruments are calibrated to minimize test
uncertainties. Furthermore, collected and stored data for each sensor was rechecked for
precision and accuracy at the end of each test scenario. The key parameters were consistently
screened to ensure tests had been performed within acceptable limits. Figure 1 is a schematic
diagram showing the locations of all sensors used to monitor test parameters.
3-55
3-56
RI (1-2-8)
e gat S
TC-1
TZ (1-4-18)
Filt ers
TC-1
TC -1
TR (1-4-1)
TC-1
TP (1-4-0)
T Q (1-4-2)
TC -1
TO (1-4-3)
DB -1
WB-1
&
W air-1
Return Air
TV (1-4-8)
TU (1-4-7)
TT (1-4-6)
TC-1
TC- 1
TC -1
TC-1
VB (1-2-26) 5f t
T BO (1-2-25) 5f t
RJ (1-2-24) 5ft
TA B (1-1-3) 3ft
TA A (1-1-2) 7f t
T S (1-4-4)
V-1
RH-1
&
DB-1
RT D-2
T BB (1-4-19) 2f t
TB A (1-4-22) 4ft
TA Z (1-4-21) 6ft
DB-1
WB -1
&
W a ri -1
RT D-1
Supply Air
DF (1-2-5)
RB (1-2-4)
T AN (1-2-3)
Return Air
@ 80F DB / 67F
WB
PB (3-2-1)
DB-1
WB-1
&
W air-1
F lexible
Duct
P s-4
P D (3-2-3)
P s-4
PA (3-2-0)
DB (3-3-5)
RD (3-3-4)
TAW (3-3-3)
TJ (3-1-13)
TI (3-1-9)
TH (3-1-8)
TG (3-1-6)
Ps-4
DC (3-3-14)
HA2 (3-3-9)
HA1 (3-3-8)
P s-4
Scale : NONE
RTD-4
DP-1
DB-1
RH-1
TC -1
T AO (3-4-29)
RPM (3-2-25)
(1-1-14)
C-1
RTD-8
DP-1
DB -1
RH 1
TC-1
TC-1
TN (3-1-15)
TM (3-1-12)
T K (3-1-14)
TA R (3-4-28)
PG (3-3-17)
SP-1
P F (3-3-16)
TBF (3-4-21)
(1-1-11)
(1-1-13)
TC- 1
P lr-1
Plr -1
TBE (3-4-22)
TC-2
mr
RTD-4
PK (3-2-5)
W com p
PF
WH (3-3-26)
W To tal
WD (3-3-6)
WB (3-2-8)
TC-1
TA S (3-4-18)
T BP (3-4-13)
P J (3-2-4)
MR (3-3-30)
TC-4
TA J (3-1-4)
TAI (3-1-3)
TAH (3-1-2)
TA G (3-1-1)
TB D (3-4-12)
HD-P-1
WCF
WA (3-2-7)
TC-1
TC-4
T L (3-1-10)
T AQ
(3-4-30)
TC-1
DD
(3-3-15)
HB2
(3-3-23)
HB 1
(3-3-22)
TA P
(3-4-31)
TC-2
R TD -2
(1-1-10)
(1-1-8)
E
v
PE
a (3-2-20)
p
o
∆P-1
r
a
t
o
r
T BH (3-4-19)
TB G (3-4-20)
R PM
W EF
WC (3-2-6)
Test Unit
RTD-1
(1-1-15)
T BN (3-4-26)
TB M (3-4-25)
TB K (3-4-23)
(1-1-12)
(1-1-9)
B (3-2-26) 5ft
VA (3-2-22) 5f t
T BQ (3-3-19) 5ft
RK (3-3-18) 5f t
TAL (3-1-11) 3ft
TF (3-4-7) 2ft
T E (3-4-6) 4ft
T D (3-4-5) 6f t
T C (3-4-4) 8f t
TB L (3-4-24)
BP-1
V -1
RH-1
&
DB-1
RTD-1
TC -4
TA F (3-4-3)
TA E (3-4-2)
TA D (3-4-1)
TA C (3-4-0)
Numbers on the right hand side indicate the quantity of each sensor
Total Channels: 131
TAK (3-1-5)
S A (3-2-21)
P C (3-2-2)
Flexible
Duct
HVAC & Monitoring Schematic Diagram
DB -1
WB-1
&
W air- 1
VD (2-1-1)
Filt ers
R H-1
TC -1
RTD-1
TB (3-1-7)
RE (3-2-19)
TA U (3-4-27)
VSD
RH -1
TC-1
RTD-1
DX Coil
Refrigeration and Thermal Test Center
Date: 05-22-2003
RTD-1
VSD
RH-1
&
DB-1
Stage
1
RG (1-3-0)
RH -1
DE (1-2-2)
RA (1-2-0)
T AM (1-2-1)
TC-4
T BC (1-1-4)
DA (3-3-2)
RC (3-3-1)
T AV (3-3-0)
V-2
V-2
VC (2-1-0)
Ultrasonic
Humidifier
TX (1-1-0)
TB I (1-2-7)
RH-1
&
DB -1
TC-1
RTD-1
RTD-1
TA X (1-1-5)
TA Y (1-4-20) 8f t
V F (2-1-3)
V E (2-1-2)
∆P-1
A ir M easurement
St ation with Pit ot Tubes
PI (2-1-5)
P H (2-1-4)
SP - Suction pressure (1)
TC - Thermocouple (39)
V - Velocit y (6)
WB - Wet-bulb temperat ure (4)
W air - Air absolute humidity (4)
W CF - Condenser fan power (1)
W com p - Compressor power (1)
W EF - Evaporator fan power (1)
W Tot al - Total system power (1)
∆P - Diff erential pressure (2)
E lectric Heater
RH (1-2-6)
T BJ (1-2-9)
2
TY (1-4-5)
ci n o s artl U
r eifi di mu H
S tag e
3
Ul trasonic
Hum idifi er
T W (1-1-1)
r et a e H ci rt c el E
li o C X D
Sensor Legend & (Quantitites) :
BP - Barometric pressure (1)
C - Condensate (1)
DB - Dry-bulb temperature (7)
DP - Dew point (2)
HD-P - Head pressure (1)
mr - Ref rigerant mass flow (1)
Plr - Liquid refrigerant pressure (2)
Ps - Statis pressure (16)
PF - Power factor (1)
RPM - Evaporator fan speed (1)
RTD - Resistance temperature device (30)
RH - Relative humidity (8)
TA (3-1-0)
RF (3-2-18)
TA T (3-4-8)
Figure 1. Detailed Test Monitoring Plan Showing Sensor Locations
Refrigerant temperatures and pressures were monitored at critical points of the
refrigeration cycle including suction and discharge of the compressor, condenser outlet,
evaporator outlet, and expansion device inlet. Air temperatures were measured at the evaporator
inlet and outlet, the condenser inlet and outlet, as well as several locations within th air
distribution system. Airflow was determined using a Wilson flow grid mounted inside the
ducting between the indoor and ambient test chambers. Condensate from the evaporator coil was
piped to a digital scale where the mass could be recorded. Room conditions such as temperature,
dew point, and air velocity were monitored in both environmental control rooms. Power demand
of the evaporator fan, condenser fan, and compressor were recorded as well. All data was
acquired through LabView 6.02 software, which included a graphical data acquisition
environment with a data logging and supervisory control add on.
Critical raw data was continuously screened for validation prior to importing the data into
the RTTC’s engineering model. The collected data points from the 20-second intervals were
averaged into one-minute intervals where necessary and used for further screening of the test
data. One-minute averages allowed data trends to be displayed with a high resolution, thus
enabling the engineering model to generate calculated hourly results such as cooling loads. The
primary data points used for comparative analysis were taken from refrigerant-enthalpy results.
Based on test design and instrumentation utilized for this project, a higher confidence was placed
on refrigerant-enthalpy data.
The one-minute averages were used to produce tabular and graphical representations of
various correlations within the engineering model. Several graphs were created to initially
screen the calculated data. After careful examination and validation of the initial screening plots,
the informational plots were produced. This final set of data plots provided relationships
between the calculated quantities. In cases where data flaws were detected, a series of diagnostic
investigations were carried out, corrections were implemented, and tests were repeated.
Test Facility
The RTTC’s HVAC testing area was used in this project. The HVAC testing area
includes both indoor and outdoor controlled environment chambers. Both chambers are served
by dedicated heating, cooling, dehumidification, humidification, and filtration systems. An
ultrasonic humidifier, controlled by a sophisticated central processing unit-based energy
management system, injected precise moisture quantities into the indoor chamber. Due to high
temperature conditions, a centrifugal humidification system was used in place of the ultrasonic
humidifier in the outdoor environment room. Each room is served by an air handling unit
equipped with split, direct expansion refrigerant coils, variable speed drive-controlled variable
air volume fans, variable supply air temperature control, and an electric heating system
controlled by pulse modulation of the power supply. The AHUs circulate air through supply and
return air plenums. A multiplex compressor rack system, consisting of two 15 horsepower (hp)
scroll compressors equipped with variable speed drive and variable suction control, serve the two
AHUs. The tested RTU was positioned inside the outdoor ambient controlled environment room
and connected to an insulated air distribution system. This air distribution system circulated
conditioned air from the test unit into the indoor controlled environment room.
Test Scenarios
All units were tested at ambient air dry-bulb temperatures of 85°F, 95°F, 105°F, 115°F,
120°F, 125°F, and 130°F.1 Air entering the evaporator coil of all units was kept at a constant
condition of 80°F dry-bulb / 60.7°F dew point, corresponding to 67°F wet-bulb, as required by
ARI 210/240.
1
The standard efficiency unit from Manufacturer C failed to operate under 130°F conditions. The high head
protection mechanism of this unit constrained the test conditions to a maximum temperature of 127°F.
3-57
Discussion of Results
Confidence Tests
Manufacturers typically publish unit performance data of their products collected at 95oF
ambient conditions, as required by ARI. Consequently, to gain confidence in test data, lab
results collected at 95°F ambient were compared with published data. The RTTC test data
yielded slightly higher gross cooling capacity than the manufacturers published data for five of
the six units (Figure 2). The average and maximum discrepancies between the manufacturer data
and test results were found to be 1,534 Btu/hr or 2.5% and 2,700 Btu/hr or 4.5 %, respectively.
It is not clear, however, whether the manufacturer’s published data used in this comparison
depends on actual ARI 210/240 tests or ARI certified simulations. Furthermore, the errors
associated with the manufacturer’s results are unknown. Hence, it is difficult to discern the exact
sources of these discrepancies.
Figure 2. Comparison of Gross Cooling Capacity from Manufacturers Published Data
and RTTC Test Data Based on 95oF Ambient Temperature
Gross Cooling Capacity (Btu/hr)
68,000
66,000
Mfg C
Mfg B
Mfg A
64,000
62,000
60,000
58,000
56,000
54,000
Standard
Hi Eff
Standard
Hi Eff
Standard
Hi Eff
Test Unit
Test Data
Mfg Data
Similar comparisons were made between test units’ and manufacturers’ compressor
power consumption data (Figure 3). The average and maximum discrepancies between the
manufacturer data and test results were 299 watts or 6.2% and 716 watts or 15%, respectively.
3-58
Figure 3. Comparison of Compressor Power Consumption from Manufacturers Published
Data and RTTC Test Data Based on 95oF Ambient Temperature
Compressor Power (kW)
7
6
Mfg B
Mfg A
Mfg C
5
4
3
2
1
0
Standard
Hi Eff
Standard
Hi Eff
Standard
Hi Eff
Test Unit
Test Data
Mfg Data
Lastly, EERs for the 95oF ambient condition were compared in the same fashion and are
depicted in Figure 4. The average and maximum discrepancies between the manufacturer data
and test results were 0.52 Btu/hr/watt or 5.2% and 0.96 Btu/hr/watt or 9.4%, respectively.
Figure 4. Comparison of System EER from Manufacturers Published Data
and RTTC Test Data Based on 95oF Ambient Temperature
System EER (Btu/hr/Watt)
14
12
Mfg A
Mfg C
Mfg B
10
8
6
4
2
0
Standard
Hi Eff
Standard
Hi Eff
Standard
Hi Eff
Test Unit
Test Data
Mfg Data
These comparisons revealed an average difference of 6.2% or less for gross cooling
capacity, system EER, and compressor power between the test results and manufacturers’
published data. As a result of these comparisons, a reasonable degree of confidence in test
design was established.
Comparison of Test Data
The main objective of this project is to quantitatively compare the operation of standard
and high efficiency units under varying ambient temperature conditions. The graphs in the
following section illustrate the quantitative test results obtained for standard and high efficiency
units of the three manufacturers. The heavier lines in all graphs represent data pertaining to the
high efficiency models.
3-59
Under ambient temperatures below 125oF the standard RTUs demanded more power than
high efficiency units (Figure 5). As temperatures increased, the power demand of high
efficiency units from manufacturers A and C grew at a much higher rate than the standard model.
As a result, the power curves of the standard and high efficiency units of manufacturers A and C
tended to converge rapidly. At an ambient temperature of approximately 130oF manufacturer
A’s high efficiency unit actually consumed 120 watts more than its standard efficiency model.
Results from Manufacturer B do not indicate any power convergence. Figure 6 intends to depict
the convergence of power use seen only in manufacturer A & C’s products. It also predicts the
power usage of manufacturer C’s standard unit at 130oF. This prediction was made by curve
fitting manufacturer C’s standard unit power usage based on experimental results as a function of
ambient temperature and extrapolating to 130oF. Figure 6 suggests that if manufacturer C’s
standard efficiency unit had operated at the 130oF ambient condition, it also may have been
surpassed by the high efficiency unit’s power usage.
Figure 5. Total System Power Consumption Based on RTTC Test Data for All Six
Standard and High Efficiency Units Subject to Various Ambient Temperatures
Total System Power (kW)
10.00
9.00
8.00
7.00
6.00
See Figure 6 for
expanded view.
5.00
4.00
80
85
90
95
100
105
110
115
120
125
130
135
o
Mfg A Std
Ambient Temperature ( F)
Mfg B Std
Mfg A Hi Eff
Mfg C Std
Mfg B Hi Eff
Mfg C Hi Eff
Figure 6. Expanded View of Figure 5 Showing the Intersection of Power Curves
at High Ambient Temperatures
Total System Power (kW)
8.00
Extrapolated
7.80
7.60
7.40
7.20
7.00
6.80
6.60
120
125
~126
~128
130
135
o
Ambient Temperature ( F)
Mfg A Std
Mfg C Std
Mfg A Hi Eff
Mfg C Hi Eff
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The increase in power usage as ambient temperature increased is attributed to the
increase in compressor work, which is a function of the compression ratio and mass of
circulating refrigerant. At high ambient temperatures, the heat rejection ability of the RTU
degrades, which results in high head pressures. Therefore, the compressor must work against a
greater pressure difference between the evaporating and condensing pressures. The high head
pressure also causes a slight rise in suction pressure. At higher suction pressures, refrigerant
becomes denser and the compressor has to compress a larger mass of vapor. Figure 7
demonstrates that, of the six units tested, only those from manufacturer A did not experience an
increase in mass flow rate at high ambient conditions. It is not clear why these units behaved
differently and this observation is being investigated.
Refrigerant Mass Flow Rate (lb/hr)
Figure 7. Refrigerant Mass Flow Rate Based on RTTC Test Data for All Six Standard
and High Efficiency Units Subject to Various Ambient Temperatures
16.5
16.0
15.5
15.0
14.5
14.0
13.5
13.0
80
85
90
95
100
105
110
115
120
125
130
135
Ambient Temperature (o F)
Mfg A Std
Mfg B Std
Mfg C Std
Mfg A Hi Eff
Mfg C Hi Eff
Mfg B Hi Eff
Each RTU’s gross cooling capacity curve dropped as the ambient temperature increased
and there was no recurring relationship between the standard and high efficiency units (Figure
8). The two manufacturer A models’ capacity decreased steadily, although the standard unit’s
performance deteriorated at a slightly higher rate than the high efficiency model.
Between ambient temperatures of 85oF and 105oF the units from manufacturer B
provided nearly equal cooling capacities. Beyond ambient temperatures of 105oF the high
efficiency unit began to outperform the standard model. The cooling capacity of manufacturer
C’s high efficiency model lagged behind standard unit when temperatures were below 100oF.
The high efficiency unit, however, maintained higher cooling capacity than the standard model
as ambient temperature increased beyond 100oF. The capacity degradation profile of both of
manufacturer A’s products stayed similar to that of manufacturer C’s standard efficiency unit.
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Gross Cooling Capacity (Btu/hr)
Figure 8. Gross Cooling Capacity Based on RTTC Refrigeration Side Test Data
for All Six Standard and High Efficiency Units Subject to Various Ambient Temperatures
70,000
68,000
66,000
64,000
62,000
60,000
58,000
56,000
54,000
52,000
50,000
48,000
80
85
90
95
100
105
110
115
120
125
130
135
o
Ambient Temperature ( F)
Mfg B Std
Mfg A Std
Mfg A Hi Eff
Mfg C Std
Mfg B Hi Eff
Mfg C Hi Eff
The energy efficiency ratio (EER) represents the relationship between net cooling
capacity and total power consumption of the units. Despite the fluctuations in cooling capacity
and power consumption, all three high efficiency units performed more efficiently than their
standard efficiency counterparts over the entire range of ambient temperatures (Figure 8).
However, as the temperature became more extreme, the EER values of the standard and high
efficiency units approached convergence.
EER Based on Refrigeration
Data (Btu/hr/watt)
Figure 8. EERs Based on RTTC Refrigeration Side Test Data for All Six Standard
and High Efficiency Units Subject to Various Ambient Temperatures
14.0
13.0
12.0
11.0
10.0
9.0
8.0
7.0
6.0
5.0
80
85
90
95
100
105
110
115
120
125
130
135
o
Ambient Temperature ( F)
Mfg A Std
Mfg B Std
Mfg C Std
Mfg A Hi Eff
Mfg B Hi Eff
Mfg C Hi Eff
Conclusions
As ambient temperatures approached extreme conditions the performance of all units’
compressors began to degrade. The compressor power demand of the high efficiency units
increased at a higher rate than that of the standard units as the ambient temperature increased.
Power consumption of the standard and high efficiency units from two of the manufacturers
began to converge as the ambient temperature increased. The power consumption and cooling
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capacity of the third manufacturer’s units showed similar and non-converging rates of
degradation. The compressor of the high efficiency unit of manufacturer A used 120 more watts
than the standard unit at the condenser inlet air temperature of 130°F. This excess power use is
greater than the uncertainty in the experiment’s power measurements. Most likely, this
phenomenon would have been repeated in manufacturer C’s products as well if its standard unit
had not shut off due to high head pressure past 127°F. Despite the convergence of power
consumption at high temperatures, the high SEER units operated at higher EERs than standard
units due to greater cooling capacity under extreme ambient conditions. Under ambient
temperatures ranging from 105oF to 125oF, higher SEER units were able to provide greater
cooling capacity while using less power than the standard units.
References
Air Conditioning and Refrigeration Institute (ARI). 1989. Standard 210/240 - Unitary Air
Conditioning and Air Source Heat Pump Equipment.
American Society of Heating, Refrigeration, and Air Conditioning Engineers (ASHRAE). 1988.
Standard 37 – Methods of Testing for Rating Unitary Air-Conditioning and Heat Pump
Equipment.
Energy Information Administration (EIA). 1999. Table C10. Electricity Consumption and
Expenditure Intensities http://www.eia.doe.gov/emeu/cbecs/pdf/alltables.pdf
Faramarzi, R. In Press. Performance Evaluation of Typical Five-Ton Roof Top Air Conditioning
Units Under High Ambient Temperatures.
Horowitz, N. 2004. Residential HVAC for Hot, Dry Climates. Natural Resources Defense
Council.
Johnson, J. 2001. New Buildings Institute Memorandum “Package HVAC Controls Initiative”.
LeRoy, J. T., Groll, E., Braun, J. E. August 1997. Capacity and Power Demand of Unitary Air
Conditioners and Heat Pumps Under Extreme Temperature and Humidity Conditions.
Ray W. Herrick Laboratories, Purdue University.
Proctor, J. E. Performance of Reduced Peak kW Air Conditioners at High Temperatures and
Typical Field Conditions. Proctor Engineering Group.
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