38 ASHRAE Jou rna l ash rae .o rg Ju l y 2005

By Michael F. Taras, Member ASHRAE

About the AuthorMichael F. Taras is manager of technology and prin-cipal staff engineer at Carrier in Syracuse, N.Y.

Is Economizer Cycle Justifi ed forAC Applications

R-410A is becoming increasingly popular in commercial air-con-

ditioning applications. Although several design enhancement

options are feasible, R-410A still exhibits a performance defi ciency

at certain environmental conditions in comparison to R-22 that is

being phased out. These issues present new challenges to system

designers and must be addressed in the most comprehensive and

cost-effective manner.

For instance, performance degradation issues for the equipment containing R-410A become more profound at relatively high ambient temperatures, where the system performance is valued and needed the most. This alone can drive technology innovation and justify its application.

On the other hand, a low-pressure refrigerant with a relatively high critical point, such as R-134a, is a viable can-

didate for R-22 substitution. However, equipment size concerns, addressed at the component level with respect to the com-pressor pump, piping and heat exchanger design, become crucial considerations and major obstacles for the implementa-tion of R-134a.

These developments have prompted the introduction of nonconventional methods to signifi cantly boost system effi ciency,

as well as incremental improvements to existing performance augmentation tech-niques. Consequently, new systems may incorporate novel design approaches, such as an economizer cycle, that in the past would not have been considered for similar applications.

The choice of a particular system design and confi guration is defi ned by essential performance characteristics, component reliability, applied cost and specifi c application considerations.

This article evaluates the advantages offered by the economizer cycle for air-conditioning and heat pump applications, and suggests that such an approach to en-hance system performance has signifi cant potential. The benefi ts of the economizer cycle are especially evident in the envi-ronment in which conventional methods

© 2005, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). Reprinted by permission from ASHRAE Journal, (Vol. 47, No. 7, July 2005). This article may not be copied nor distributed in either paper or digital form without ASHRAE’s permission.

Ju l y 2005 ASHRAE Jou rna l 39

reach a plateau of rapidly diminishing returns on investment or become physically impossible.

Economizer CycleThe economizer cycle is well-justifi ed for high compression

ratio conditions, such as refrigeration applications, where the 30% increase in performance provides a substantial advantage. Until recently, the economizer cycle was not considered a cost-effective solution for air-conditioning installations where conventional technology was pushed to the limit. Oversized heat exchangers (condensers and evaporators) reached their ef-fectiveness plateaus and compressor volumetric and isentropic effi ciencies approached the respective maximums.

Now, however, the implementation of environmentally sound refrigerants has negatively affected system performance at certain environmental conditions, and industry standards and govern-ment regulations have raised the bar for the minimum effi ciencies of the air-conditioning equipment, a trend that will continue.

This course of events has opened the door to various kinds of alternate technologies, encouraging a more serious evaluation of the economizer cycle.

Principle of OperationAlthough the economizer cycle schematics may differ with

respect to specifi c system confi guration and implementation details, two basic design solutions exist and are shown in Figures 1 and 2.

The economizer cycle is more complex than the conventional system, integrates additional components, and entails more sophisticated control algorithms. When the economizer cycle is in operation, a portion of liquid high-pressure refrigerant is diverted from the main circuit and rerouted through the econo-mizer loop, where its temperature and pressure are reduced to some intermediate level in an auxiliary expansion device. Therefore, the colder economizer fl ow can be used, through the heat transfer interaction in the economizer heat exchanger, to further boost subcooling of the main refrigerant. This extra subcooling allows for the enthalpy difference increase in the evaporator and subsequent system performance augmentation (see the P-h diagram in Figure 3).

Downstream of the economizer heat exchanger, the auxiliary fl ow is returned to the intermediate compressor port, usually at the superheated state controlled by the auxiliary expansion device. When the colder economizer fl ow is mixed with the partially compressed suction vapor, temperature of the latter is reduced, benefi tting the compression process. Furthermore, the bypass loop between the economizer and suction compressor ports, also exhibited in Figures 1 and 2and 2and , can be used for the

Until recently, the economizer cycle

was not considered a cost-effective

solution for air-conditioning installa-

tions where conventional technology

was pushed to the limit. Oversized

heat exchangers (condensers and

evaporators) reached their effective-

ness plateaus and compressor volu-

metric and isentropic effi ciencies ap-

proached the respective maximums.

40 ASHRAE Jou rna l ash rae .o rg Ju l y 2005

Hence, it is not uncommon for the economizer heat exchanger effective-ness to reach 90%, promoting extremely close temperature approaches of nearly 1°F (0.6°C).

The compressor injection scheme is of high signifi cance as well. Therefore, the intermediate compression pockets must have an optimal and restricted exposure to the economizer ports, limiting cyclic parasitic compression and expansion losses associated with the pressure dif-ference in the economizer loop and in the aforementioned compression pockets. For instance, it was observed that the economizer effectiveness reduction by 10% to 15% or use of the conventional compressor injection scheme leads to a decline in the economizer cycle perfor-mance gains by 30% to 50%, making the approach extremely diffi cult to justify.

PerformanceARI Standard Rating Conditions.

As discussed earlier, the two strategies, namely equal capacity approach and equal effi ciency tactic, are feasible. In each scenario, it is prudent to determine the performance enhancement minimum for the economizer cycle, particularly at the ARI standard rating conditions* for commercial equipment (ARI Standard 340/360). A midsize packaged scroll

compressor system of premium efficiency would serve as a logical contender for such a study.

For the purpose of this analy-sis, the condenser and evapo-rator dimensions and circuit-ing were preserved, although circuiting optimization could provide a performance en-hancement potential.

The test results closely matched by the analytical

model predictions are summarized in Table 1. The R-410A economizer cycle provides a performance boost of at least 4% to 5% at the ARI standard rating point. Also, it becomes obvious that the economizer cycle may not be required to match the R-22 performance for some premium effi ciency systems.

However, if further performance enhancement is desirable or a number of systems of various capacities share an identical

system unloading strategy. (As known, the fl ash tank, separating liquid and va-por phases and essentially representing a 100% effective heat exchanger, can be used in place of the economizer.)

Although the economizer cycle ad-vantages are obvious, the full benefi ts of the concept are slightly diminished by three phenomena concurrently oc-curring in the system. First, the evapo-ration temperature is slightly reduced, negatively impacting refrigerant fl ow rate at the compressor suction port and subsequently system capacity. This side effect of extra subcooling is of second-order signifi cance and, since it equally affects capacity and power, no change in the system effi ciency is registered.

Second, the refrigerant fl ow rate in the condenser is increased, promoting condensation temperature elevation. Since the economizer fl ow is relatively small in air-conditioning applications, this concern can be easily addressed through proper condenser design.

Last, additional compressor power consumption is expected, as extra economizer fl ow must be compressed from the intermediate to the discharge pressure. These three phenomena moder-ate, but do not completely overshadow, the performance benefi ts achieved in the “ideal” economizer cycle, at most operating conditions.

The economizer cycle per-formance diminishes with the pressure ratio and subsequent decrease of the available thermal potential in the economizer heat exchanger, approaching the con-ventional system characteristics at the point where the aforemen-tioned losses become equal to the performance gains.

The economizer cycle advan-tages can be used in two ways. First, an equivalently sized economized compressor can be used to deliver more capacity to the system. In another approach, the compressor size reduction by about 8% to 10% leads to a system effi ciency boost, while preserving capacity character-istics at the Air-Conditioning and Refrigeration Institute (ARI) standard rating point.

Managing all system losses, especially in the compressor and economizer heat exchanger, are crucial for the economizer approach to be successful in air-conditioning applications.

*The ARI standard rating conditions are 80°F (26.7°C) dry bulb/67°F (19.4°C) wet-bulb indoor air temperatures, and 95°F (35°C) outdoor air temperature.

Outdoor Air

2 1

3 4 5

Indoor Air

Evaporator

Bypass Valve

Economized Compressor

77

Condenser

66Economizer Expansion

Device

Economizer HX

Main Expansion

Device

Figure 2: Economizer cycle (alternate).

Outdoor Air

12

Economizer HX

6

Economizer Expansion

Device

Main Expansion

DeviceDevice

3 4 5Main 3 4 5Main

Indoor Air

Evaporator

Bypass Valve

Condenser

77Economized Compressor

Figure 1: Economizer cycle.

Critical Point

3 2 1

66 6 7

77

77

4 5

Figure 3: P-h diagram for economizer cycle.

P

h

Ju l y 2005 ASHRAE Jou rna l 41

75 80 85 90 95 100 105 110 115 120 125 13075 80 85 90 95 100 105 110 115 120 125 130Ambient Temperature, °F

1.05

1

0.95

0.9

0.85Cap

acit

y R

atio

(Rel

ativ

e to

R-2

2)

ARI Rating Point

R-410A Economized

R-410A Conventional

R-134a Conventional

75 80 85 90 95 100 105 110 115 120 125 130Ambient Temperature, °F

1.05

1

0.95

0.9

0.85

EER

Rat

io (R

elat

ive

to R

-22)

ARI Rating Point

R-410A Economized

R-410A Conventional

R-134a Conventional

Figure 4: Minimum capacity degradation. Figure 5: Minimum EER degradation.

Table 1 (left): Minimum economizer cycle benefi ts at ARI standard rating conditions. Table 2 (right): Minimum performance degradation at a high ambient temperature of 125°F (51.7°C).

Design Identical Reduced Option Compressor Size Compressor Size

Capacity EER Capacity EER

Design Basis (R-22) 100% 100% 100% 100%

R-410A Economized 105% 100% 100% 104%

R-410A Conventional 100% 100% 95% 104%

Design 95°F (35°C) 125°F (51.7°C) Option Outdoor Temp. Outdoor Temp.

Capacity EER Capacity EER

Design Basis (R-22) 100% 100% 100% 100%

R-410A Economized 100% 104% 95% 95%

R-410A Conventional 100% 100% 91% 88%

chassis size, the economizer cycle becomes a useful technique to achieve a desired target. In such cases, twice as much en-hancement (in comparison to the numbers exhibited in Table 1) can be obtained from the economizer approach, while the conventional technology will experience rapidly diminishing returns on investment or become physically impossible.

One of the side benefi ts of the economizer cycle is its aug-mented dehumidifi cation capability, due to lower evaporation temperatures, that potentially can be used in hot and humid environments to replace the reheat coil methodology.

High Ambient. When a life-cycle cost analysis of any air-con-ditioning equipment is performed, system operation at a wide spectrum of environmental conditions must be examined.

It is not unusual that a high ambient temperature region at-tracts particular attention, since at those conditions, the system performance is most often demanded and valued by end custom-ers. Any refrigerant system suffers performance degradation in high ambient environments, particularly nowadays when the operation requirements are stretched to temperatures as high as 135°F (57.2°C).

The downward trend for the equipment containing the R-410A refrigerant is more profound than the respective decline for the R-22 systems. In some cases, the performance parity at the full-load ARI standard rating conditions (and at the 80°F/67°F-80°F [26.7°C/19.4°C-26.7°C] part-load conditions) is not a suffi cient criterion to obtain identical life-cycle cost characteristics over the entire system operation envelope.

Distinct design points of the equivalent R-410A performance must be chosen to adequately represent a region of high ambient temperatures for various classes of applications. Some R-410A systems must have superior performance at the ARI conditions as compared to the equivalent R-22 equipment, to have parity at high ambient temperatures.

The packaged commercial unit of premium effi ciency that we used in this study serves here as well to determine mini-mum performance degradation for the R-410A systems at high ambient temperatures.

Table 2 displays performance characteristics of the R-410A systems as compared to the equivalent R-22 systems operating at the performance parity at the ARI standard rating condi-tions. The 125°F (51.7°C) ambient temperature limit refl ects the absolute minimum demanded by a majority of commercial applications today.

The R-410A economizer cycle reveals superior performance and is lagging the R-22 system by only 5%, while the conven-tional R-410A system exhibits twice as much performance degradation. Obviously, the gap between the R-410A and R-22 conventional systems (as well as the R-410A conventional and economizer systems) becomes much deeper for less effi cient equipment and at higher ambient temperatures. As a result, the equipment will accumulate more hours at less effi cient opera-tion to satisfy external load demands at those conditions.

For R-134a, similar critical temperatures and somewhat analogous shapes of the two-phase region in the vicinity of the critical point suggest that both R-134a and R-22 refrigerants exhibit comparable performance degradation at high ambient temperature conditions. Furthermore, the R-134a refrigerant has a lower compressor discharge temperature (by 25°F to 30°F [14°C to 17°C]), promoting higher reliability and translating into a wider compressor operation envelope.

The relative performance results for all three refrigerants are presented in Figures 4 and 5and 5and , with the R-134a indeed revealing less performance reduction than the R-410A at high ambient temperatures.

Last, we can conclude that it requires signifi cantly less effort and investment to reach performance parity at high ambient

42 ASHRAE Jou rna l ash rae .o rg Ju l y 2005

for the R-410A economizer systems, while any conventional approach may become thermodynamically prohibitive.

Part-Load Operation. A conventional multicircuit system unloads by turning circuits on and switching them off to satisfy external load demands and to control temperature and humidity within the conditioned space.

For the economizer system, multiple additional steps of un-loading can be accomplished naturally, reducing life-cycle cost of equipment and augmenting part-load performance charac-teristics. More particularly, the economizer cycle capacity can be decreased by bypassing a portion of the refrigerant between the economizer and suction compressor ports (Figures 1 the economizer and suction compressor ports (Figures 1 the economizer and suction compressor ports ( and2) as well as by operating the system in the economizer and conventional modes.

This approach offers four operation modes for each inde-pendent circuit within the system: 1) conventional mode, 2) economizer mode, 3) conventional mode with bypass, and 4) economizer mode with bypass.

These four modes offer many opportunities for the unloading strategy. Further-more, adjustable fl ow control devices can be used for the economizer expansion and bypass valves, offering an infi nite number of unloading steps through modulation or pulsation techniques and extra fl exibility in control of vari-ous operational parameters.

Such parameters may in-clude (but are not limited to) system capacity, compressor discharge temperature and power, and conditioned space temperature and humidity. For instance, for abnormally high ambient temperatures, a sequence of small unloading steps can limit compressor power while preserving system performance, extending the operation envelope, and preventing nuisance shutdowns.

Thus, continually changing load demands are satisfi ed with greater precision, keeping the conditioned space parameters within the comfort zone, eliminating temperature and humid-ity variations, and enhancing system reliability and effi ciency through a reduction of start-stop cycles.

Manufacturing Cost. Cost is one of the most critical ingredi-ents in validation and justifi cation of nontraditional technology, especially at the initial stages of its application. The assump-tion is that presently manufactured equipment containing HFC refrigerants is produced and sold at a premium. (This trend may change with the refrigerant and oil price correction due to a boost in production volumes of the HFC refrigerants and phaseout of R-22.) On the other hand, the newly introduced systems must operate at least at the existing effi ciency levels to adequately satisfy customer expectations and perception. There-fore, a lower boundary for the system performance is established

and a decision is made regarding incremental design augmenta-tions and the appropriate methodology to be employed. This decision is based on a cost performance tradeoff.

The economizer cycle offers signifi cant advantages in sys-tem operation and performance, but it becomes particularly diffi cult to justify this concept from a cost standpoint for the commercial systems below 5 tons (17.6 kW) in capacity, since this equipment is extremely cost sensitive. Consequently, it is expected that the conventional techniques of performance enhancement will prevail in the market segment of small com-mercial packaged equipment while alternate technology will become affordable for the larger systems.

Obviously, it is especially diffi cult to establish a hard uniform cost basis across various platforms and product lines, due to variability in refrigerant-oil and material prices, inconsistency in chassis sizes and compressor frames, disparity in coil manu-facturing capability and dependence on particular manufacturer labor and burden rates. The numbers presented in this section

should be viewed only as ra-tional estimates.

The fi rst step is to assess the costs associated with convert-ing larger systems (above 5 tons [17.6 kW] in capacity) to the HFC refrigerants while preserving the effi ciency lev-els. To reach a general conclu-sion independent of system size and confi guration specif-ics, the analysis was conducted on a relative scale with the normalization to the R-22 standard system cost.

It was observed that for the R-410A systems this incremental cost is between 5% and 10%, and results primarily from the pre-miums paid for R-410A components that can achieve effi ciency levels comparable to the existing systems. The lower values are associated with the smaller equipment sizes (5 to 15 tons [17.6 to 52.7 kW]), since these systems usually have relatively large cabinet structure and, hence, operate at lower condensation temperatures (further away from the critical point).

The R-134a equipment exhibits 12% to 16% cost increase with signifi cantly smaller variation with respect to the equip-ment size, which can be explained by substantially higher critical point for this refrigerant and lower sensitivity to the condensation process.

Extra costs associated with the economizer cycle are related to the complexity of manufacturing an economized compressor, an economizer heat exchanger, an auxiliary expansion device, and economizer loop piping, plus the cost of the extra refriger-ant charge. Obviously, cost becomes less of an issue when it is physically impossible for the traditional technology to achieve the desired effi ciency levels.

For the reasons outlined earlier, the economizer cycle becomes more advantageous for the larger equipment,

30

25

20

15

10

5

0

Cos

t P

rem

ium

, Per

cent

9 9.5 10 10.5 11 11.5 12 12.5Energy Effi ciency Ratio, Btu/h · W

Standard 90.1-1999

20 Ton System

15 Ton System

Figure 6: Effi ciency cost premium (R-410A systems).

Ju l y 2005 ASHRAE Jou rna l 43

but a cost parity point is quite different for R-410A and R-134a refrigerants.

Once again, to eliminate relevance to a particular equipment type, the study was carried out on a percentage basis with respect to a traditional R-410A system cost. For the R-410A systems, the economizer cycle cost becomes 3% to 5% lower than the conven-tional unit cost for the equipment capacity range of 18 tons (63.2 kW) and above, comparable for the 10 to 15 tonnage (35.1 to 52.7 kW) range and somewhat higher below 10 tons (35.1 kW). For the R-134a refrigerant systems in the 15 to 20 tonnage (52.7 to 70.3 kW) range, the economizer cycle cost is already 6% lower than the conventional unit, alluding to the fact that this concept has a wider spectrum of applications with R-134a refrigerant. This phenomenon can be explained simply by the fact that the pressure ratios for the R-134a refrigerant are higher in general, further promoting the economizer cycle approach.

Equipment cost increases exponentially with the effi ciency level, therefore, it is essential to evaluate the R-410A system cost trends. It is logical to an-ticipate that the normalized cost curves for the units of various sizes would collapse into a single trend line, but the analysis suggests just the opposite. For instance, two representative cost curves for the systems of 15 and 20 tons (52.7 to 70.3 kW) in capacity are exhibited in Figure 6, where the least expensive method for each incremental effi ciency step is plotted. Although both costs are normalized by the respective system cost values, the growth rate for the 20 ton (70.3 kW) system is signifi cantly higher than for the 15 ton (52.7 kW) system.

The explanation is related to two phenomena. First, there is a somewhat disproportional cost increase for the system compo-nents, such as compressors and coils, relative to the increase in the equipment size. This disproportionate cost increase is due to limited manufacturing capability and inadequate competi-tion in the market.

Second, the larger equipment tends to operate closer to a critical point because of a relatively small cabinet struc-ture, and therefore, requires an extra step for performance augmentation.

The results described previously assume effi cient manufac-turing operations and can be extended to a wider spectrum of equipment sizes, although it is expected that the difference in cost will be smaller for the lower capacity equipment and just the opposite will be true for the larger systems.

Other Performance-Boosting ContendersAlthough the compressors with integrated liquid injec-

tion ports are becoming readily available on the market, the

economized confi gurations are still limited, particularly for air-conditioning applications.

This shortage of the economized compressors for air-conditioning applications can be attributed mainly to their design requirements and the extreme diffi culty of manag-ing the losses associated with the economizer ports. Thus, original equipment manufacturers have started to search for non-conventional thermodynamically similar arrangements using traditional compression technology.

Such options may include (but are not limited to) natural refrigerants, secondary circuits for subcooling enhance-ment, cascade systems, entropy recovery expansion devices, minichannel (brazed aluminum) heat exchangers as con-densers and evaporators, liquid-suction heat exchangers, and economizer-like designs. The last two options have close resemblance to the economizer cycle and are worth further consideration.

Designs incorporating a liquid-suction heat exchanger are

well known in the refrigeration industry and provide the most advantages for the refrigerants with high

characteristics. This concept, shown in Figure 7Figure 7Fi , works reason-ably well for the refrigeration applications, where a thermal potential between the refrigerant vapor exiting the evapora-tor and liquid leaving the condenser is signifi cant enough to provide adequate heat transfer interaction between these two refrigerant streams. As a result, an extra subcooling of liquid refrigerant entering an expansion device appreciably enhances evaporator performance. With air-conditioning equipment, the aforementioned temperature difference is substantial-ly reduced.

This smaller temperature difference, along with a lower vapor density entering the compressor suction port, makes the performance benefi ts and this system concept practicality questionable.

Outdoor Air

Condenser

Compressor

Liquid-Suction HX

Evaporator

Indoor AirExpansion

Device

Figure 7 (left): Cycle with liquid-suction heat exchanger. Figure 8 (right): Cycle with auxiliary heat exchanger (Confi guration 1).

Auxiliary HX

Auxiliary Expansion

Device

Outdoor Air

Condenser

Compressor

Evaporator

Indoor AirIndoor AirExpansion

DeviceDevice

44 ASHRAE Jou rna l ash rae .o rg Ju l y 2005

Next, we will consider the economiz-er-like schematics. Four arrangements are available based on the location of the splitting and mixing points of the auxiliary and main refrigerant fl ows.

Figure 8 shows the splitting point up-stream of the auxiliary heat exchanger, while the mixing point is positioned at the compressor suction. An alternate arrangement (not shown) would locate the splitting point downstream of the auxiliary heat exchanger.

Figure 9 exhibits the splitting point downstream of the auxiliary heat exchanger, while the mixing point is placed downstream of the evaporator. An alternate arrangement (not shown) would locate the splitting point upstream of the auxiliary heat exchanger.

The proposed technique is based on traditional compression technology and an addition of the auxiliary heat exchanger (preferably in the counterfl ow arrangement) to the system schematic. This design approach provides extra subcooling and overall system performance boost by establishing an optimal refrigerant fl ow fraction that bypasses the evaporator to achieve extra subcooling in the auxiliary heat exchanger.

Although various design confi gurations are feasible, the refrigerant fl ow entering the evaporator is also proportionally reduced in all such schematics. As a result, the economizer-like schematics offer only a fraction of performance enhancement, in comparison to the economizer cycle.

Consequently, the economizer cycle still stands out as the most superior and feasible option in the category.

Heat Pump ApplicationsHeat pumps simultaneously serve cooling and heating markets,

covering broader applications and growing in importance. Such a trend offers a tremendous opportunity for the economizer cycle, since the heat pump equipment usually operates at the elevated pressure ratios (in comparison to the traditional air-conditioning in-stallations), while having suffi ciently high performance targets.

The economizer cycle has only recently entered the air-con-ditioning arena, and most likely will penetrate the heat pump market in the near future. Heat pump designs incorporating an economizer loop are more complex than the traditional systems and, respectively, include an additional four-way valve, an ad-ditional economizer heat exchanger or an extra main expansion device–check valve assembly. Obviously, many variations of these basic designs (primarily related to the economizer loop connections and insertion of additional components such as an accumulator) are foreseeable.

Although extra system complexity, additional hardware, and increased cost are to be justifi ed by the benefi ts obtained from the economizer cycle performance, a vacuum in potential contenders will further promote the technology.

Concluding RemarksIt is apparent that the economizer

cycle offers many benefi ts for air-con-ditioning applications and stands out among other potential contenders, but also introduces several design, cost, and application challenges that must be properly addressed. In particular, the economizer cycle provides a competi-tive advantage, where other methodolo-gies fail to be economically viable or become physically impossible.

A typical performance augmentation for the economizer system is expected to be approximately 8%. Such an im-provement can easily compensate for the R-410A performance defi ciency at

the ARI standard rating point for commercial equipment. Furthermore, the economizer cycle demonstrates its superior-

ity at high ambient operation, where it outperforms conventional systems by 10% on average. Although this is a signifi cant improvement, issues associated with R-410A performance degradation in such environments must be addressed further, potentially by a combination of the economizer cycle and tra-ditional technology.

Other benefi ts of the economizer cycle are related to fl exibility in the unloading strategy, improved dehumidifi cation capability and superior operation in the heat pump installations.

The economizer concept can be used in a multicircuit system confi guration as well, where economizer heat exchangers as-sociated with each circuit can be combined into a single unit, offering a reduction in the heat exchanger cost by 25%.

As mentioned earlier, all the economizer cycle advantages can be realized only if special attention is paid to the econo-mized compressor injection scheme and the economizer heat exchanger effectiveness, each of which may easily reduce foreseeable benefi ts by a factor of two.

Additionally, it should be expected that the economizer sys-tem control logic will be more complex, the number of compres-sor suppliers will be limited, and the development cycle will be longer, at least until suffi cient experience is gained.

Although, in principle, any system can use the economizer cycle, today the concept most likely is not economically viable for systems below 10 tons (35.1 kW) in capacity. However, for the larger systems, the economizer cycle cannot be matched by any other realistic contenders at a comparable cost.

Although strong arguments can be presented, based on perfor-mance, for R-134a to be the refrigerant of choice for packaged commercial equipment, the cost concerns develop into a major obstacle, even for the economizer cycle.

BibliographyLemmon, E.W., et al. 2002. NIST Reference Fluid Thermody-

namic and Transport Properties – REFPROP 7.0. National Institute of Standards and Technology at www.nist.gov/srd/nist23.htm

Outdoor Air

Indoor Air

Evaporator

Expansion Device

Auxiliary Expansion

Device

Auxiliary HX

Compressor

Condenser

Figure 9: Cycle with auxiliary heat exchanger (Confi guration 2).