A S H R A E J O U R N A L a s h r a e . o r g S E P T E M B E R 2 0 166 8

Daniel H. Nall, P.E., FAIA, is vice president at Syska Hennessy Group, New York.

COLUMN ENGINEER’S NOTEBOOK

Daniel H. Nall

Direct-expansion (DX) cooling and heat pump systems have often been the less-respected alternative for building conditioning compared with hydronic systems. In general, they are considered to be a “cheaper” system, with higher maintenance cost, shorter lifespan, limited controllability and lower energy efficiency. On the other hand, they are often more convenient, more amenable to design-build procurement, and less demanding of operator expertise.

Recent improvements, however, to commercially

available products have made these systems more con-

trollable, and more energy efficient, to the point that

they are competitive with chilled-water systems, both

from the standpoint of preciseness of environmental

control and energy efficiency.

This column will describe an office building project in

New Jersey for which custom direct-expansion air con-

ditioning was more energy efficient, and less costly both

to construct and to operate than a built-up chilled-water

system. Since this building was finished, several manu-

facturers have included some of the components incor-

porated in the custom rooftop systems of the project as

options in their standard product lines.

IntroductionDX systems of old were the bottom of the ladder

for environmental control. Originally equipped with

reciprocating compressors, the units were loud, inef-

ficient and had limited controllability. An article from

ASHRAE Journal in 2001 discussed means of evaluating

dehumidification limitations for DX units,1 primarily

addressing the loss of humidity control when the unit

cycles for part-load and humid outdoor air is supplied

continuously. Unloading issues with DX systems often

left various forms of reheat as the only means of accurate

dry-bulb temperature and humidity control within the

space.

Engineers sought various methods of integral reheat

using the refrigeration cycle itself to raise the tempera-

ture of the supply air to reduce sensible capacity while

avoiding energy expenditure for reheat and somewhat

increasing the efficiency of the refrigeration process.2

Even before continuously unloading centrifugal water

chillers were available, chilled-water systems were

able to provide a relatively constant temperature cool-

ing source, without thrashing the stepped unloading

devices, because the thermal mass of the circulating

water dampens the impact of stepped unloading. Even

with slightly varying chilled-water temperature, vari-

able volume air-handling units can, energy efficiently,

vary sensible cooling to meet space loads while main-

taining an acceptable range of the apparatus dew point

to control space humidity. The advent of the centrifugal

chiller, furthermore, not only facilitated part-load cool-

ing control, but also increased the efficiency of cooling

production far beyond the level achievable by DX sys-

tems using positive displacement compressors.

BY DANIEL H. NALL, P.E., BEMP, HBDP, FAIA, FELLOW/LIFE MEMBER ASHRAE

Direct Expansion or Chilled Water–Which is Better?

©ASHRAE www.ashrae.org. Used with permission from ASHRAE Journal at www.syska.com. This article may not be copied nor distributed in either paper or digital form without ASHRAE’s permission. For more information about ASHRAE, visit www.ashrae.org.

S E P T E M B E R 2 0 16 a s h r a e . o r g A S H R A E J O U R N A L 6 9

COLUMN ENGINEER’S NOTEBOOK

is located, while providing energy efficiency superior to

a built-up chilled-water system.

The Chilled-Water AlternativeChilled water systems could be considered the “gold

standard” for high performance building cooling sys-

tems. They have excellent part-load performance,

continuous unloading, flexibility of connection to ter-

minal units to convey cooling to occupied space, and

ease of control. In general, water-cooled chiller plants

are acknowledged as the most energy-efficient cool-

ing source for large buildings. For this study, a chiller

plant complying with ASHRAE Standard 90.1-2013 was

assumed, using the higher EER and lower IPLV option

for Path B, effective Jan. 1, 2015. Following are the code

requirements for the chilled-water plant:

• 0.585 kW/ton (0.166 kW/kW) – chiller power for

units ≥400 tons (1407 kW) and ≤600 tons (2110 kW)3;

• 0.052 kW/ton (0.0148 kW/kW) – cooling tower fans

(40.2 gpm/hp [3.4 L/s·kW], 2.8 gpm/ton [0.05 L/s·kW]

condenser water)4;

• 0.053 kW/ton (0.015 kW/kW) – condenser water

pumps (19 W/gpm [0.084 W/L·s], 10°F range [5.6°C])5;

• 0.044 kW/ton (0.013 kW/kW) – chilled-water pumps

(22 W/gpm [0.097 W/L·s], 12°F range [6.7°C])5;

• 0.73 kW/ton (0.21 kW/kW) – cooling plant energy;

• 20.5 Btu/Wh (6.01 J/J) – compressor-only EER

(6.01 COP); and

• 16.3 Btu/Wh (4.8 J/J) – cooling plant EER. (4.79 COP).

Adding an ASHRAE Standard 90.1-2013-compliant fan

yields the following power requirements:

• 80°F/67°F (26.7°C/19.4°C) entering dry-bulb/wet-

bulb temperature;

• 0.0015 hp/cfm (0.0024 kW/L·s) fan power;

• 0.34 kW/ton (0.10 kW/kW) of cooling fan power;

• 1.07 kW/ton (0.304 kW/kW) total HVAC power; and

• 11.2 Btu/Wh (3.28 J/J) – EER for the system

(3.29 COP).

Comparison of the two cooling source alternatives

required standardization of the air-handling sec-

tion of the system. After rigorous energy modeling

studies, a cold (50°F [10°C]) air delivery strategy was

adopted, because the reduced fan energy more than

offset the additional cooling energy required for the

lower supply air temperature. Because the reduced

mixed airflow has a higher than normal outdoor air

fraction, the mixed air condition onto the coil has a

Current DX systems can be competitive with chilled-

water systems both for control accuracy and energy effi-

ciency due to two technical advances. The first of these is

the ability of small compressors to perform continuous,

rather than stepped, unloading. Prior to the advent of

this technology, small compressors typically had discrete

unloading steps, which for DX units translated directly

into variations in supply air temperature and the appa-

ratus dew-point temperature.

Continuous unloading of small compressors can be

achieved either by variable speed drive, or by timed

rapid cycling of the compressor between a loaded and

unloaded state. The time ratio between loaded and

unloaded states determines the part-load fraction of the

compressor, while the cycling frequency is so fast that

the refrigerant pulses blend into a steady stream, albeit

at a reduced flow rate.

The second innovation is the electronic expansion

valve. This innovation allows more precise handling of

varying refrigerant flow rates, to allow control of evapo-

ration temperature in the face of varying coil loads. With

these two innovations in place and with the increas-

ing efficiency of small compressors, DX systems can

approach the precision and efficiency of chilled-water

systems.

While most packaged DX systems are air-cooled, sev-

eral manufacturers are offering evaporative condensers

as an option for their premium semi-custom packaged

rooftop systems. This measure typically increases the

efficiency of the condenser and compressor section of

the unit by more than 30%. For the larger versions of

such systems, the ASHRAE Standard 90.1-2013 perfor-

mance requirement is for a minimum EER of 11.7 with

an IEER of 11.9.3 This performance requirement com-

pares favorably with the cumulative ASHRAE Standard

90.1-2013 requirements for all-air VAV systems with

water-cooled chillers as the source of cooling.

For a recent build-to-suit office building project in

New Jersey, with aspirations of Platinum Certification

for both LEED NC v.3 and LEED CI v3, a number of HVAC

options were considered and subjected to rigorous

energy modeling and cost estimating study. As a result

of these efforts, a custom packaged direct-expansion

rooftop system with evaporative condenser was recom-

mended and used on the project. The resulting appli-

cation retains many of the advantages of the packaged

rooftop units used in the office park where the building

A S H R A E J O U R N A L a s h r a e . o r g S E P T E M B E R 2 0 167 0

• 0.98 kW/ton (0.28 kW/kW) total HVAC power; and

• 12.3 Btu/Wh (3.60 J/J) – EER for the system

(3.60 COP).

The part-load performance of the chilled-water

plant mandated by ASHRAE Standard 90.1-2013, Path

B, requires an IPLV of 0.380 kW/ton (0.108 kW/kW).

Typically, this path is for centrifugal compressors using

variable speed drive for capacity control.

Prefabricated Evaporative Condenser Dx PenthouseThe system selected for comparison with the chilled-

water plant option was a custom packaged rooftop pent-

house, two of which were used to condition the building.

The two systems had the following characteristics:

• Evaporative condenser with VFD fans;

• Sump temperature-controlled circulating pumps;

• Three variable speed magnetic bearing compres-

sors, one with hot gas bypass, total capacity 405 tons

(1424 kW);

• Nominal compressor power 0.54 kW/ton

(0.15 kW/kW);

significantly higher enthalpy than the AHRI rating

condition.

As a result, fan power per ton (kW) of refrigeration is sig-

nificantly less than the standard condition. However, meet-

ing the supply air temperature goal, given the incoming air

condition, with 44°F (6.7°C) chilled water requires either

a very low coil face velocity, 425 fpm (2.2 m/s), or a coil that

does not meet ASHRAE Standard 62.1-2013 standards for

cleanability (8 rows, 12 fpi). For this study, the impact of the

very close approach of coil leaving air temperature to enter-

ing chilled-water temperature has been ignored. Below are

the specifications for the actual air side of the system and

the resultant EER for the system:

• 84°F/69°F (28.9°C/20.6°C) – entering dry-bulb/wet-

bulb temperature;

• 50°F/50°F (10.0°C/10.0°C) – supply air dry-bulb/

wet-bulb temperature;

• 85,500 cfm (40 714 L/s) – supply and return air

volume;

• 126.9 fan bhp (94.6 kW) – total fan horsepower;

• 0.24 kW/ton (0.07 kW/kW) of cooling fan power;

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A S H R A E J O U R N A L a s h r a e . o r g S E P T E M B E R 2 0 1672

• 0.068 kW/ton (0.019 kW/kW) – condenser fan;

• 0.012 kW/ton (0.003 kW/kW) – circulating pump;

• 0.62 kW/ton (0.18 kW/kW – cooling plant energy;

• 22.2 Btu/Wh (6.50 J/J) – compressor only EER

(6.50 COP); and

• 19.4 Btu/Wh (5.68 J/J) – cooling plant EER

(5.68 COP).

The airside system is considered to be the same perfor-

mance as for the chilled-water system. The actual DX coil

furnished for the system was 8 row, 10 fpi, with a face veloc-

ity of 440 fpm (2.24 m/s), so it is very near the pressure

drop of the coil required for the chilled-water system:

• 84°F/69°F (28.9°C/20.6°C) – entering dry-bulb/wet-

bulb temperature;

• 50°F/50°F (10.0°C/10.0°C) – supply air dry-bulb/

wet-bulb temperature;

• 85,500 cfm (40 714 L/s) – supply and return airflow

• 126.9 fan bhp (94.6 kW) – total fan horsepower;

• 0.24 kW/ton (0.07 kW/kW) of cooling – fan power;

• 0.86 kW/ton (0.25 kW/kW) – total HVAC power; and

• 14.0 Btu/Wh (4.10 J/J) – EER for the system (4.10 COP).

• Integrated supply fan array with twenty-one

16 in. (406 mm) airfoil centrifugal fans, total airflow

85,500 cfm (40,714 L/s); and

• Integrated return fan array with fourteen 20 in.

(508 mm) airfoil centrifugal fans.

Despite the fact that air is supplied at a low tempera-

ture (50°F [10°C]), and is coming onto the coil at a higher

enthalpy than normal, using a direct-expansion coil

with a low face velocity allows this low-temperature

air to be generated with a refrigerant saturated suction

temperature of 44.2°F (6.8°C). A supply air temperature

reset algorithm is installed to maximize airside econo-

mizer utilization for the system.

The higher saturated suction temperature allowed by

the direct-expansion coil results in an efficiency of the

small centrifugal chillers in the packaged system that

approaches or exceeds that of the larger compressors in

the built-up chilled-water plant. Below are the perfor-

mance parameters for the cooling supply section of the

submitted custom packaged DX system:

• 0.54 kW/ton (0.15 kW/kW) – compressor power;

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A S H R A E J O U R N A L a s h r a e . o r g S E P T E M B E R 2 0 1674

FIGURE 1 Custom penthouse plan.implications for the

occupant’s exit strategy.

A building located in a

large office park, totally

occupied by buildings

with packaged rooftop

units, that requires

additional operating

staff expertise may

be at a disadvantage

when offered for sale or

lease in the real-estate

marketplace.

Studying the pack-

aged option through

the use of a competi-

tive design-assist pro-

curement enabled the

owner to select a very

cost-effective and high

performance alterna-

tive. Table 1 shows a

comparison of the two

systems at full and part

loads, showing that the packaged DX system maintains

its energy efficiency advantage even at low part loads.

The final advantage of the packaged rooftop system

was the reduction in construction cost. Two of these cus-

tom packaged DX rooftop systems were purchased at a

savings of $1,367,300 compared with the chilled-water

plant option for this 292,000 ft2 (27 138 m2) building.

Figure 1 is a plan of the final packaged rooftop design.

While many of the energy-efficiency alternatives included

in this unit were only available at that time in fully cus-

tom packaged units, several options, including inverter-

driven compressors and evaporative condensers, are

now available as standard options on premium packaged

rooftop lines. These new units are competitive with chiller

water systems both for first cost and operating efficiency.

References1. Doty, S. 2001. “Applying DX equipment in humid climates.”

ASHRAE Journal 43(3):30 – 33.2. Taras, M. F. 2004. “Reheat: Which concept is best.” ASHRAE

Journal 46(12):34 – 39.3. ANSI/ASHRAE/IES Standard 90.1-2013, Energy Standard for Build-

ings Except Low-Rise Residential Buildings, Table 6.8.1-1.4. ANSI/ASHRAE/IES Standard 90.1-2013, Energy Standard for Build-

ings Except Low-Rise Residential Buildings, Table 6.8.1-3.5. ANSI/ASHRAE/IES Standard 90.1-2013, Energy Standard for Build-

ings Except Low-Rise Residential Buildings, Table 6.8.1-7.6. ANSI/ASHRAE/IES Standard 90.1-2013, Energy Standard for Build-

ings Except Low-Rise Residential Buildings, Table 11.5.2-1, Note e.

TABLE 1 Full load, part-load and IPLV performance for chilled-water system vs. custom DX penthouse.

CH ILLED WATER ASHRAE STANDARD 90.1-2013AIRSIDE

APPLICATIONDX–

ACTUAL

PART LOAD

CH ILLER (W/TON)

CT FAN (W/TON)

CW PUMP

(W/TON)

CHW PUMP

(W/TON)

FAN (W/TON)

EER EER

100% 585 52 53 44 243 12.3 14.075% 380 39 53 25 176 17.8 18.650% 374 35 24 16 137 20.5 20.925% 386 26 24 11 109 21.6 15.9

IPLV 380 19.4 19.3

ConclusionThe initial reason for the consideration of the custom

packaged direct-expansion air-conditioning units as

an alternative to a built-up chilled-water plant had to

do with the facility staffing for the occupant. A chilled-

water plant would require personnel with more exten-

sive training than would a packaged system, even a cus-

tom packaged system.

Not only would that personnel requirement have budget-

ary implications for the occupant, but it also would have

COLUMN ENGINEER’S NOTEBOOK