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Improvement of IEER Rating and Dehumidification Capability in Unitary DX Equipment.


This paper addresses a cost-effective improvement of energy efficiency and dehumidification performance of unitary DX air-conditioners, an equipment type that cools roughly three-quarters of commercial and two-thirds of government building floor space. The choice of a particular DX package unit design and configuration is defined by essential performance characteristics and component reliability, along with market penetration and cost constraints. This paper evaluates the advantages offered by application of a simple refrigeration cycle variant, and suggests that such an approach to enhance system performance has significant potential for cost-effective reduction of operating energy.

The benefits of the modified cycle are particularly apparent as traditional methods for raising efficiency reach a plateau of rapidly diminishing return on cost or become physically infeasible. Oversized heat exchangers (condensers and evaporators) are near their effectiveness limits, and conventional compressor volumetric and isentropic efficiencies have approached their respective theoretical maximums. In addition, refrigerant changeover from HCFC-22 to HFC-410a, has negatively affected system performance at high ambient temperatures.

Building codes, driven by ASHRAE and AHRI, utility incentives, government policy, and market demand have raised the bar for the minimum efficiencies of DX air-conditioning equipment, a trend that will undoubtedly continue. Increased attention to dehumidification stems from reduced sensible loads, along with demands for improved comfort and acceptable indoor air quality, while reducing energy use. This requires equipment to match the desired sensible-to-latent load ratio, not only at rated conditions, but also over the full range of off-design conditions. Higher R-value building envelopes reduce sensible load, so latent loads are now a larger proportion of the total cooling load. This course of events has opened the door to various kinds of alternate technologies, and encourages a more serious evaluation of refrigerant cycle variations.

Cooling equipment that can efficiently meet latent load independent of sensible load will realize the energy savings potential of improved building designs. Conventional unitary air-conditioning equipment has a sensible heat ratio (SHR) that is typically 0.70 to 0.75 at standard AHRI rating conditions. In many operating scenarios, conventional equipment latent capacity is inadequate for space dehumidification demands. Since most air-conditioning units are controlled by thermostats responding only to sensible loads, the result is excess humidity in the building space or application of energy-intensive reheat. This situation is exacerbated by condensed water that remains on the evaporator coil after compressors cycle off and re-evaporates, since the blower often runs constantly to meet ventilation requirements. ASHRAE has recognized these realities by including paragraph 5.9 in Standard 62.1-2010. Paragraph 5.9 states that systems shall also be analyzed so that occupied spaces shall have a maximum 65%RH when the design dewpoint is used and the solar load is zero. These have the effect of lowering the cooling load SHR by 20 to 30 points (-0.20 to -0.30).


Optimizing heat transfer in the evaporator coil is a balance between physical coil size, air-side differential static pressure, and refrigerant-side pressure loss and evaporation temperature. The refrigerant in a conventional cycle evaporator coil is fully evaporated well upstream of the coil exit, mostly to protect the compressor from damage due to ingesting incompressible liquid. Analysis using the DOE/ORNL Heat Pump Design Model and NIST CYCLE_D identified that refrigerant condition through the evaporator coil can be controlled to increase phase change heat transfer by maximizing liquid refrigerant fraction throughout the coil length.

Maximizing the presence of liquid refrigerant at the boiling point in the evaporator coil significantly increases EER (Energy Efficiency Ratio, Btuh cooling per Watt electric power) because boiling heat transfer coefficient in the liquid-nucleate regime is much greater than in vapor-mist refrigerant regime, where droplets boil in suspension and the tube walls are dry. Vapor temperature increases as it absorbs heat, while evaporating liquid does not, which results in a colder bulk coil temperature and increased latent capacity. The authors' cycle variant (U.S. patent 6,427,454) achieves increased liquid fraction at the inlet and the outlet of the evaporator coil, a reduced mist regime, and minimized vapor heat transfer. At the coil inlet, refrigerant sub-cooling to near the evaporator saturation temperature results in halving the vapor fraction at the exit of the expansion device. At the coil outlet, refrigerant flows to a high-effectiveness heat exchanger/accumulator, which vaporizes the liquid before it enters the compressor.

In the fully developed turbulent flow regime for pure liquid or for vapor at Re above 10,000, the most widely accepted correlation for the Nusselt number according to the Wolverine Heat Transfer Data Book is the Sieder-Tate correlation:

[Nu.sub.D] = 0.027[Re.sup.4/5.sub.D][Pr.sup.1/3][([mu]/[[mu].sub.s]).sup.0.14]

where the heat transfer coefficient is found by solving for h


In the two-phase flow regime before the mist point the Boyko-Kruzhilin with the Reynolds number of the pure liquid flow:

[Nu.sub.mik] = 0.024 * [Re.sup.0.8.sub.lo][Pr.sup.0.43.sub.1][([Pr.sub.1]/[Pr.sub.1,w]).sup.0.25]

where the two-phase heat transfer coefficient is found by solving for h

[mathematical expression not reproducible]

The mist point is the distance along the tube where boiling transitions from nucleate to suspended droplets and is calculated using the Lockhart-Martinelli Parameter.

[mathematical expression not reproducible]


[m.sub.l] is the liquid phase mass flowrate and [m.sub.g] is the gas phase mass flowrate.


Perhaps most significantly, the added ability to manage refrigerant conditions bestowed by the cycle variation releases constraints on other equipment design parameters. Additional subcooling provides an enthalpy difference increase in the evaporator and subsequent system performance augmentation. The enthalpy difference increase combined with the higher average refrigerant heat transfer coefficient allows selection of an evaporator coil for higher face velocity and reduced face area, which tends to lower cost. The reduced face area combined with the ability to tolerate liquid at the evaporator coil exit allows evaporator bypass of return air, which greatly improves latent performance as needed. Further gains are realized by allowing part of the air to bypass the evaporator through a modulating damper that dynamically varies the latent capacity to match the load sensible heat ratio--the unit becomes a variable sensible heat ratio machine with a fully protected compressor inlet along with an increase in energy efficiency. With higher bypass fractions, the thermal potential between the refrigerant vapor exiting the evaporator and liquid leaving the condenser is increased, which further enhances evaporator performance.

Refrigerant heat transfer coefficients were calculated for various coil geometries, refrigerant mass flows, phases, subcooling and circuiting. As refrigerant evaporates, the flow velocity increases along the length of the tubes because the vapor is much less dense. These range from 2,000 to 7,000 Btuh/[ft.sup.2]/deg-F (10 to 40 W/[m.sup.2]K) in the nucleate boiling regime. This is 14 to 70 times greater than the mist and pure vapor phase heat transfer coefficients ranging from 100 to 400 Btuh/[ft.sup.2]/deg-F (600 to 2300 W/[m.sup.2]K). Modeling a nominal 5-ton (17.5 kW) circuit with an R-410a mass flow of 150 lb/hr/ton (240 kg/hr/kW) through a coil with 1 feed per ton (1 feed per 3.5 kW) and an average tube velocity of 10 ft/sec (3 m/s), here is how the heat transfer compares: (a) In a standard cycle refrigerant nucleate boiling occurs along 54% of the coil tube length, mist boiling along 26%, and superheat along 20% of the coil, giving an average refrigerant heat transfer coefficient of 1374 Btuh/[ft.sup.2]/deg-F (7802 W/[m.sup.2]K); (b) In the subcooled zero-superheat cycle variant, nucleate boiling occurs along 68% of the coil and mist boiling along 32%, giving an effective heat transfer coefficient of 1698 Btuh/[ft.sup.2]/deg-F (9642 W/[m.sup.2]K). The mixture at the exit of the expansion device is 19% vapor by mass, or 71% vapor by volume; (c) And thirdly, in the subcooled separated cycle variant, nucleate boiling occurs along 96% of the coil, predicting a coefficient of 1955 Btuh/[ft.sup.2]/deg-F (11,101 W/[m.sup.2]K). This is a 26% and a 42%, respectively, increase in average refrigerant heat transfer coefficient.


Analysis of a standard commercially available R-410a scroll compressor using NIST CYCLE_D software at 115 F (46.1 C) condensing temperature and 45 F (7.2 C) evaporating temperature yields the following results. The p-h state diagrams are shown in Figure 2. NIST CYCLE_D was used to isolate the effect of thermodynamic refrigeration cycle variations, so the calculated improvement in efficiency is exclusive of gains from the increase in heat transfer effectiveness, air-side modifications and changes in unit operation. The standard cycle COP is 3.41 / EER 11.6. The variant cycle COP of 3.74 / EER 12.8 shows an improvement of 9.6% assuming no liquid and zero superheat at the evaporator exit. COP is improved further by allowing liquid at the evaporator exit; however, the software does not yet have that capability.

Detailed modeling of two commercial package units using the DOE/ORNL Heat Pump Design Model Mark 7.04a provides the following results. This DOE/ORNL-developed software is a rigorous and highly detailed research tool for use in the steadystate design analysis of unitary equipment. It is based on state of the art air-side and refrigerant-side heat transfer and pressure drop correlations, and AHRI compressor maps that have been verified through ASHRAE 1173-TRP. The software was used to model four operating conditions for two types of equipment, each both in standard factory configuration and with the cycle variant and component changes applied. This modeling takes into account all of the effects of thermodynamic refrigeration cycle variations, thus the calculated improvement in efficiency includes secondary offsets as well as synergistic gains from the increase in heat transfer effectiveness and equipment modifications.

The four operating conditions satisfy the full load and part-load IEER rating criteria set forth in ANSI/AHRI Standard 340/360-2007, and the IEER (Integrated Energy Efficiency Ratio weighted Btuh per Watt) of each unit configuration was calculated accordingly. Equipment types that were modeled are an R-410a dual-circuit 8 1/2-ton (29.8 kW) 13.2 IEER TXV-equipped package unit, and an R-22 dual-circuit 20-ton (70.1 kW) 11.2 IPLV orifice-equipped rooftop unit. Figure 3 shows model output for the R410a standard factory configuration. Figure 4 compares the performance results given by the modeling for the R-410a unit (model-C) and the R-22 unit (model-T).

Results show an IEER increase of 19.7% for the modified R-22 unit and 27.1% for the modified R-410a unit, relative to the standard factory rated configuration for these units. It is expected that refrigerant 410a will better respond to liquid-suction heat exchange due to its higher Cp[DELTA]T/[h.sub.fg] characteristics at the variant operating points. The cycle variation combined with a change of coil circuiting from face split to interlaced in model-C increases part-load dehumidification capacity along with contributing to the greater cooling capacity and IEER increase than obtained with the model-T.


The variant cycle IEER improvement can be realized any of three ways with the application of innovative system controls: (1a) an equivalently sized compressor can be used to deliver more capacity with the ability to vary SHR to match dehumidification load as needed, or (1b) an equivalently sized compressor can be used to deliver more latent capacity to the system without a loss of efficiency as needed, possibly eliminating the need for reheat, or (2) the compressor size can be reduced to further boost efficiency, while preserving capacity characteristics at the AHRI standard rating point. If a number of models of different capacities share an identical chassis size, the variant cycle could be a useful technique to achieve desired performance targets. For example, results for model-C show 10-tons (35 kW) of cooling capacity at standard conditions with a unit rated for 8.5 nominal tons (29.8 kW) prior to modifying the cycle. In such cases, the variant cycle can be particularly cost effective, while conventional technology such as increasing coil size experience rapidly diminishing returns on investment or become physically unfeasible.

Cost is one of the most critical factors in valuation and justification of alternate technology. Added costs associated with application of the variant cycle and equipment modifications are related to the liquid suction heat exchanger accumulator, bypass damper, controls, copper piping, and additional refrigerant charge. Figure 5 gives an example of cost premium for the variant cycle in a 10-ton (35 kW) unit. The calculated 16% manufactured cost premium for the additional components is partially offset by use of either a reduced size evaporator coil or a reduced capacity compressor. This cost premium can be economically recovered through reduced operational energy costs.


The variant cycle as presented offers a competitive advantage, where other methodologies of further raising IEER ratings fail to be economically viable in the cost-sensitive light-commercial market, are impractical, or incur performance degradation. An important benefit of the cycle variant and equipment modifications is augmented dehumidification capability, due to lower evaporation temperatures, that could be used in humid climate zones to replace the reheat coil methodology, giving further energy savings and equipment cost reduction.


Thanks to the many people who helped make this work possible, including Dr. Jim Galvin and Dr. Jeffery Marqusee of the DoD Environmental Security Technology Certification Program (ESTCP), Ted Cherubin and Michael Taras of Carrier Corporation, Kevin Riley of Indyne Corporation, and Christopher Cook of Booze Allen Hamilton.


Phil Baggett, John Murphy, Paul Solberg, and Justin Wieman. "Energy-Saving Strategies for Rooftop VAV Systems" Trane, Engineers Newsletter Live, 2010.

Michael West, Ph.D, P.E. "Low-Cost Dedicated Outdoor-Air Systems" HPAC Engineering, July 2010.

John Dieckmann, Kurtis McKenney, and James Brodrick, Ph.D. "Vapor Compression, Cromer Cycle: Energy-Efficient Dehumidification" ASHRAE Journal, August 2009.

Donald P. Gatley, P.E., and Jason LeRoy. "Old Dehumidification Technology: Liquid Subcooling/Air Reheating" ASHRAE Journal, December 2009.

Rice, Keith, Zhiming Gao, and Bill Jackson. Development of DOE/ORNL Heat Pump Design Model Mark 7 Version, Oak Ridge National Lab, February 2006.

Michael F. Taras. "Is Economizer Cycle Justified for AC Applications" ASHRAE Journal, July 2005.

Michael F. Taras. "Is Reheat: Which Concept is Best" ASHRAE Journal, December 2004.

Detlef Westphalen,Ph.D., Kurt W. Roth, Ph.D., John Dieckmann, and James Brodrick, Ph.D. "Improving Latent Performance" ASHRAE Journal, August 2004.

Brown, J. S.; Domanski, P. A.; Lemmon, E. W. CYCLE D: NIST Vapor Compression Cycle Design Program. Version 5.0

Detlef Westphalen, Ph.D. "New Approach to Energy Savings For Rooftop AC" ASHRAE Journal, 2004.

Karim Amrane, Ph.D., Glenn C. Hourahan, and Greg Potts. "Latent Performance of Unitary Equipment" ASHRAE Journal, January 2003

Stanley A. Mumma, Ph.D., P.E., and Kurt M. Shank. "Achieving Dry Outside Air in an Energy-Efficient Manner"

ASHRAE Journal, 2001.


IEER -- Integrated Energy Efficiency weighted Btuh per Watt according to ANSI/AHRI Standard 340/360-2007

DX -- Direct Expansion

DOE/ORNL -- U.S. Department of Energy/Oak Ridge National Laboratory

EER -- Energy Efficiency Ratio

HCFC -- Hydrochlorofluorocarbon

HFC -- Hydrofluorocarbon

ASHRAE -- American Society of Heating, Refrigeration and Air Conditioning Engineers

AHRI -- Air-Conditioning, Heating and Refrigeration Institute

R-value -- Resistance to heat flow, used in insulation

SHR -- Sensible Heat Ratio

%RH - % Relative Humidity

NIST CYCLE_D -- National Institute of Science and Technology Vapor Compression Cycle Design software

COP -- Coefficient of Performance

p-h -- pressure enthalpy diagram

TXV -- thermal expansion valve

IPLV -- Integrated Part Load Value

FPM -- feet per minute

CFM -- cubic feet per minute

LS-HXGR -- liquid suction heat exchanger accumulator

Michael K. West, Ph.D., P.E.


Thomas Brooke, P.E.


Michael West, Ph.D, P.E. is Principal-CEO at Advantek Consulting Engineering, Melbourne, Florida. Tom Brooke, M.B.A., C.E.M., P.E. is Senior HVAC Engineer at Advantek Consulting Engineering, Ocala, Florida.
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Author:West, Michael K.; Brooke, Thomas
Publication:ASHRAE Conference Papers
Date:Jun 22, 2013
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