R744 - THERMODYNAMIC POSSIBILITIES AND ...r744.com/assets/link/Hrnjak_THERMODYNAMIC POSSIBILITIES...- 3 - 2010 International Symposium on Next-generation Air Conditioning and Refrigeration

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2010 International Symposium on Next-generation Air Conditioning and Refrigeration Technology, 17 – 19 February 2010, Tokyo, Japan

THERMODYNAMIC POSSIBILITIES AND TECHNOLOGICAL

OPPORTUNITIES FOR IMPROVING THE SCIENCE OF REFRIGERATION IN SEARCH FOR LOW GLOBAL WARMING

THERMAL SYSTEMS

P. Hrnjak,

Res. Professor, Dep. of Mechanical Science and Engineering, The University of Illinois, Urbana, USA Abstract:

The request of new low GWP options based on either synthetic (man made) refrigerants or natural (naturally occurring) materials have to offer high efficiency in addition to environmental friendliness and safety for people and property. That situation will open new possibilities for drop in refrigerants along options for natural refrigerants. The paper will present some options from the point of thermodynamic limits and opportunities for new technologies to reach those limits and/or affect some of perceived constraints and discuss experience from immediate past. It is good to remember the recent history from the point of thermodynamic limits and opportunities for new technologies to reach those limits and/or affect some of perceived constraints. This paper presents an overview of expectations as we begin to reconsider CO2 systems versus those using R134a, HFO1234yf and other synthetic refrigerants and discusses the effects of thermodynamic properties on cycle performance and thermophysical considerations on the system. A brief overview of the results of the Alternative Refrigerant Cooperative Research Program (CRP) of SAE International—probably the most comprehensive and fair experimental comparative analysis for mobile air-conditioning (MAC) systems—is used as an illustration of the approach and provides real results. Paper also present the influence of the development of CO2 technologies on R134a and other systems, i.e., transfer of technology. Finally, paper presents the progression of MAC system performance with the clear statement that the differences among refrigerants in same-size systems were significantly smaller than the improvement of all systems indicating that similar situation is to be expected to continue. Regardless of the outcome, the beneficiaries of the competition are the entire professional community and humanity as a whole.

Key Words: Refrigerants, Evaluation, System 1 INTRODUCTION The last two decades are characterized with various pressures on synthetic refrigerants from the point of their environmental effects. The first wave was due to ozone depletion when excellent-performing and easy-to-work-with CFCs were replaced with HFCs. The recent pressure is due to the global warming impact of HFCs. A similar situation as in the case of CFCs is now being repeated: two major applications are under stronger pressures than others mostly due to significant refrigerant leaks: automotive air conditioning and commercial refrigeration. Other stationary applications are following the trend. It is generally accepted that automotive air conditioning contributes the most to direct emissions (refrigerant leak) due to the extremely large number of open compressor units (on the order of hundreds of millions) and harsh vibrating conditions. Commercial refrigeration systems have significantly fewer number of units in operation (nearly three orders of magnitude fewer rack systems compared to automotive systems) but each unit contains a refrigerant charge in the range of several hundred kilograms to few tons (three orders of magnitude more). Leak detection is difficult. Refrigerant leak rates of 40% of charge per year seen as realistic at the end of 20th century are reduced by better design and maintenance to leak rates of order of 10-15% of the system charge (with disputes about exact number). The situation concerning the effect of refrigerant in mobile air conditioning systems on global warming in Europe was the most pressing because it is difficult to follow the obligations from the Kyoto Protocol in a segment of the industry (automotive air conditioning) that

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experienced booming expansion in the period the Kyoto Protocol applies. Other industry groups, mostly in the food and beverage segment, have reacted to public pressure, including several companies with strong name recognition. After evaluating several options for an environmentally friendly refrigerant (refrigeration system) in both segments (automotive a/c and commercial refrigeration) carbon dioxide had emerged as the chosen alternative to currently used synthetic refrigerants. We had witnessed the rise of several synthetic refrigerant options with a GWP less than 150 (but which are flammable and sometimes have uncertain toxicity). The question was: will effect of reduced cradle to grave direct effect (essentially leak of refrigerant and generation in production) be able to offset assumed lower efficiency of carbon dioxide systems (thus higher indirect emission) and assumed higher cost? Most of the disputes were (and for some still are) in assumed lower thermodynamic efficiency of CO2 cycle. Many debates were intense and frequently biased for various reasons. Nevertheless, even among scientists polarization was evident. While most of the analytical, modeling studies were indicating that it was impossible to make a more efficient system with CO2 than with the R134a typically used in the automotive and commercial sectors, several experimentalists were providing results that indicated better performance of the system with CO2. Discussions were mostly along the lines of necessary simplifications of the models and the impossibility to get identical sizes and technologies of equipment when doing experiments. CO2 system designers had to use several new approaches and even technologies that were not used in R134a systems at that time: internal heat exchangers, microchannel heat exchangers (evaporators in particular), and eventually an ejector as the most realistic expansion-work-recovery device. Many of these technologies originally used in CO2 systems are now adopted and commercially used in MAC R134a systems. In fact, these technologies are being commercialized earlier in R134a systems than in CO2 systems simply because there are still no commercialized CO2 MAC systems yet. Company Denso provides an excellent example: they already commercialized for R134a microchannel evaporators, internal heat exchangers and ejectors (although internal heat exchangers and ejectors in R134a systems offer far less advantage than in R744 systems). It is still unclear why there was insufficient collaboration between modelers and experimentalists or why models were not made for the systems that were used in experiments to better understand both experimental results and gage the models. The confidentiality of some of the designs and even test data were probably an inhibiting factor. Probably the most fair and comprehensive experimental programs were several Alternative Refrigerant Cooperative Research Programs (CRP) of SAE International. We will present some of the most important findings from that program later with intention to remind us of possible similar scenario for search for stationary applications. There are still several open issues when discussing comparisons of various technologies or refrigerants. One is: is it time to move from discussions about the fairness of comparison between refrigerants and focus on real systems (i.e., short term market issues)? Another is: What is the limit for increasing efficiency and how do possibilities for various refrigerants differ (i.e., the question of a long term viability)? How can components be optimized to take maximal advantage of the thermophysical properties? How can we add components that are both economically viable and increase the efficiency of the thermodynamic cycle or the system? Maybe less strong drive for global solution in stationary applications. 2 WHAT WAS EXPECTED Any simplified analysis (either cycle or even decent system model) of the performance (COP) of the synthetic refrigerants (here R134a) that almost exclusively operate in subcritical mode show lower a COP for R744 up to 40-50%. So, interest in carbon dioxide as a refrigerant did not come from a desire to realize potentials in performance (as may be the case for ammonia or propane) but to explore possibilities of an otherwise good candidate for a working fluid: namely, CO2 is naturally occurring, nonflammable, and nontoxic (although not without any potential problem). Thus, the interest was to explore the potentials and reasons for the differences compared to conventional fluids and perhaps to find ways to reduce them. One of the first good analyses that indicated the potential of CO2 came

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from J. Pettersen, 1999 and most of the elements of their approach are still valid. Figure 1 illustrates the shortcomings of the CO2 cycle on a T-s diagram. Basic problems are in the heat rejection in the supercritical mode (transcritical operation) where, because of significantly higher temperature differences, cycle efficiency is reduced due to increased entropy generation. Moreover, there is a higher compression work as well as expansion losses in isenthalpic process without work recovery when compared to subcriticaly operating fluids (like R134a in this example). Of course, it is clear that the Rankine or Evans-Perkins cycle is not the best framework for R744. Any improvement that would:

1. Reduce discharge temperature: multistage compression, intercooling, and more isothermal-like compression, and

2. Reduce losses on the “left side” such as use of expanders, ejectors, internal heat exchangers, external cooling that reduce the throttling loss and similar

will benefit CO2 significantly more than conventional refrigerants operating in the subcritical regime. Internal heat exchangers were used from the very beginning while work recovery options (ejector, expander or something else) came later. Some manufacturers were enthusiastic about two-stage compression while others are still looking at other options.

Fig. 1. Comparison of CO2 (green) and R134a (red) in simple Evans-Perkins cycle

The elements that were mostly neglected in the early simplified investigations were the effects of thermophysical properties on heat transfer in both heat exchangers. An excellent overview of the thermophysical properties of CO2 is presented in a paper by Kim, Pettersen and Bullard, 2003. Excellent thermal conductivity combined with extremely high specific heat, especially in the zones of the highest importance (exit from the gas cooler), in addition to the absence of subcooling in the gas cooler results in an unexpectedly low approaching temperature differences (ATD) at the gas cooler outlet. High discharge temperatures, a significant drawback in cycle efficiency, are an advantage in heat exchanger performance. The results of these differences are schematically presented in Fig. 2.

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Length

Tem

pera

ture

CO2

HFC

Air

LMTD

ATD

Length

Tem

pera

ture

CO2

HFC

Air

LMTD

ATD

Fig. 2. Typical profiles in counter flow heat rejection of CO2 (green) and R134a (red). ATD stands for Approaching Temperature Difference

3 WHAT WAS LEARNED IN ALTERNATIVE REFRIGERANT

COOPERATIVE RESEARCH PROGRAMS OF SAE INTERNATIONAL Prior to the comprehensive research and evaluation program made under the umbrella of SAE International, a complete project was conducted at ACRC of the University of Illinois with the identical objective. In a very careful experimental program, two mobile air conditioning systems were compared: an off the shelf R134a a/c system of the worldwide present compact car was compared to a prototype 744 system (see Fig. 4).

V = 4.32 lFace area 1964 cm2

A air = 7.2 m2

A ref = 0.40 m2


A air = 5.2 m2

A ref = 0.49 m2

R744gas cooler

R134acondenser

R744V=3.71 l

Face area 408 cm2

A air = 4.2 m2

A ref = 0.66 m2

R134aV=3.72 l

Face area 405 cm2

A air = 3.5 m2

A ref = 0.55 m2

R134a

V=155cm3

Swashplate

R744

V=20.7cm3

R134a

R744

Compressor

Oilseparator

Internal HX

Condenser

AccumulatorEvaporator

Gas cooler


A air = 7.2 m2

A ref = 0.40 m2


A air = 5.2 m2

A ref = 0.49 m2

R744gas cooler

R134acondenser

R744V=3.71 l

Face area 408 cm2

A air = 4.2 m2

A ref = 0.66 m2

R134aV=3.72 l

Face area 405 cm2

A air = 3.5 m2

A ref = 0.55 m2

R134a

V=155cm3

Swashplate

R744

V=20.7cm3

R134a

V=155cm3

Swashplate

R744

V=20.7cm3

R134a

R744

Compressor

Oilseparator

Internal HX

Condenser


Gas cooler

R134a

R744

Compressor

Oilseparator

Internal HX

Condenser


Gas cooler

Figure 4 Schematic drawings of the system and photographs of components of two systems in early studies of

potential for R744 in MACs

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The systems and procedure were described in detail in papers by Yin et al., 1998, Boewe et. al 1999. In essence, the prototype R744 system was designed in such way that the heat exchanger face areas were the same and the volume of the heat exchangers did not exceed the volume of the baseline R134a HXs. Air flow rates were identical and pressure drops on the air side were also the same (for all practical purposes). The R744 system was allowed to have an internal heat exchanger (IHX). IHXs were not used at that time in R134a MAC systems, and the dominant assumption in the industry that IHXs were not feasible for R134a. Both systems were explored under the same conditions over a wide range of temperatures and humidities. The major results are shown in the Fig. 5. One of the consequences of using off-the-shelf components for the baseline R134a system was that the technology and design of the baseline R134a components and system were optimized for cost rather than for the performance. On the other hand, all designs and sizes of the new R744 system were constructed to fit every aspect of the baseline (frontal areas, arrangements, even capacity), which complicated the design process, increased the cost, and virtually ensured that the CO2 system could not take full advantage of fluid properties in, especially having in mind the limitations of resources and knowledge in a given timeframe.

Figure 5. The cooling capacity of the two systems were kept constant while efficiencies show differences for two

compressor speeds as a function of ambient temperatures while all other conditions remain constant

SAE International provided a framework for a significant (multi-year and multi-million dollar) project funded by almost every company in the industry. The execution was at ACRC and CTS (Creative Thermal Solutions). The technical part of the project was almost identical to the earlier ACRC project. The main advantage was the significant support of all the participants to the technologies they were interested in. Strict confidentiality was imposed to protect confidentiality and motivate participating companies to provide the best components known to them. Only outer dimensions of equipment and generic descriptions were revealed. Results were released to the participants immediately and to the public one year after the final presentation. The project execution was monitored on an almost daily basis by the Expert team. The Expert Team was comprised of representatives (excellent engineers) of four major OEMs and had identified leaders for each technology (refrigerant, system). Members of the Expert Team, as well as the participating companies had very different initial inclinations, but all the issues were resolved in open and constructive discussions. The Expert Team and the Executive Team frequently reported results to all participants. All systems had the same outer sizes of heat exchangers, the same air flow rate and the same test conditions. Refrigerant-side designs were allowed to be adjusted to each of the systems to the best of the knowledge of the engineers involved under the management of the Expert team member who was in charge of that particular system. The capacity of all systems were adjusted to be the same so that the only measure of difference was COP, at least in principle. To understand potentials, a separate set of conditions were run to evaluate capacity at the same operating conditions, or even maximizing capacity to understand the tradeoffs with COP. The project had several phases:

Phase 1. Evaluation of the baseline R134a system, R744 system, “enhanced” R134a system (EnR134a) and propane in a secondary loop with 50% Ethylene Glycol/water mixture. In addition, a 2002 technology R744 system was explored that had an evaporator of the same

10 15 20 25 30 35 40 45 50 55 600

1

2

3

4

5

21.1/2.83

26.7/4.96

32.2/7.08

3000rpm,Vc=35.4m3/min

1800rpm, Vc=26.9m3/min

950rpm,Vc=22.67m3/min, R134a Steady state, Cycling , R744 Steady state, Cycling

Teai(oC)/Ve(m3/min)

Cap

acity

[kW

]

Condenser Air Inlet Temperature [oC]

10 15 20 25 30 35 40 45 50 55 600.0

0.5

1.0

1.5

2.0

2.5

3.0

3.5

4.0

4.5

5.0

21.1/2.83

26.7/4.96

32.2/7.08

3000rpm,Vc=35.4m3/min

1800rpm, Vc=26.9m3/min

950rpm,Vc=22.67m3/min, R134a Steady state, Cycling , R744 Steady state, Cycling

Teai(oC)/Ve(m3/min)

CO

P

Condenser Air Inlet Temperature [oC]

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technology as the “enhanced” R134a system but thinner sized and with an overall counterflow gas cooler. The findings indicated that the baseline system was very good with a variable displacement compressor and sizeable condenser, but the “enhanced” system had even better components, mostly HXs. This obvious advantage in HX technology was the main reason for the project extension to the 2002 R744 system that had the same new evaporator technology (same manufacturer) as the EnR134a system, new counter flow gas cooler, and smaller compressor displacement (to help reduce inefficiencies in very destroked conditions). Phase 2. Evaluation of a more compact baseline R134a system and an equally sized R152a direct expansion (DX) system. The major objective was to explore R152a (as a low GWP fluid) and compare it as a drop-in replacement for R134a in existing systems. The major conclusion was that R152a has better potential for higher capacity and efficiency, but also has flammability issues when used as a drop-in replacement. IMAC 30/50 (Improved Mobile Air Conditioning). To address concerns that the R134a systems in previous phases were optimized for cost and reliability, this program was focused on efficiency improvements of the R134a system only: several IHXs were tried, condenser improvements were examined, expanders and ejectors were discussed, new compressor types were explored developed specifically for R134a, etc. The goal was 30% higher COP and 50% lower refrigerant leakage and it was achieved. Nevertheless, compressor efficiency of the best R134a compressors was still lower than R744 compressors in Phase 1. CRP-150. This phase was focused on synthetic alternative refrigerants. Several teams were focused on various aspects: toxicity, materials compatibility, performance, servicing, etc. Four refrigerants were evaluated by the performance team: R134a (as a baseline), “Fluid H”, “AC-1” and “Solvay” in dual and single evaporator systems. Results indicated close performance, slightly lower than R134a, but all were abandoned due to toxicity issues that surfaced. CRP-1234yf. It was focused on this new refrigerant as alternative to R134a. This refrigerant was the only one left in the arena at that time after all the previously mentioned refrigerants, in addition to” DP1” and “JDH” were abandoned. Results indicated good performance, a bit lower than the baseline R134a.

For our purpose, the most important conclusion is that from the Phase 1 stabilized efficiency and capacity tests. The Expert team concluded that for system capacity higher ambient conditions are most important for occupant comfort so the systems are compared at maximum capacity at ambient temperatures 45ºC, 35ºC and 25ºC (see Table 1). For annual energy analysis the moderate and lower ambient temperatures are more important because they have higher weighting factors. COPs are presented at equal capacity for comparison at 35ºC, 25ºC and 15ºC in Table 2.

Table 1 System capacities [kW] at MAX capacity Table 2 System COPs at EQUAL capacity

Green fields represent the maximal achieved values in each category. Obviously, CO2 systems were very good in providing high cooling capacity at any condition while efficiencies for the 2002 R744 system were better at lower ambient temperatures and higher speeds. Better performance at lower ambient temperatures is a consequence of the cycle (thermodynamic properties) because at these conditions the R744 system operates in subcritical mode. R744 systems outperformed R134a MAC systems at higher compressor speeds because of the significant effect of refrigerant pressure drop through the HXs. Due to the required high compactness of the MAC R134a systems, evaporators and condensers are significantly smaller and have higher refrigerant side pressure drop for the same compressor displacement compared to stationary applications. High pressure fluids (like R744 here or R410A) are much less COP sensitive to pressure drops than low-pressure R134a due to p-T relationship (thermophysical properties) because temperature differences due to pressure drops affect

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cycle efficiency. At some conditions in the R134a systems temperature drops due to refrigerant flow through the evaporator reach 10ºC.

Similar information is presented in the Fig. 6. It shows efficiencies of the 2002 R744 system vs. the EnR134a system for the same capacity. The 2002 R744 system had the same COP as EnR134a system at 25oC in idling (900 rpm) while at driving condition (2500 rpm) equal COP occurred at 35oC. Based on much longer operating hours at lower ambient temperatures CO2 proponents predict lesser total energy consumption while R134a (and new low GWP fluid proponents) point out lower first cost and simplicity of the systems. Most likely a few years ago this positioning of both sides would have been surprising.

Figure 6. Efficiencies of 2002 R744 (green) and EnR134a (red) systems for the same capacity based on results

in SAE AR CRP project, Phase 1

4 THE LIMITS OF COP

It has been demonstrated that thanks to good thermophysical properties and design, MAC R744 systems are for all practical purposes equal performers as R134a. The logical question is what will happen when a call for energy efficiency pushes both system to increasing HX surfaces or improved components? Or: would all of these conclusions from MAC system extend to RAC (stationary) systems where weight and space are less important constraints. Or, yet another way: what are the limits of the increasing COP trend and how do fluids properties affect it. Certainly, a thermodynamic analysis of the cycle provides the first answer. Nevertheless the real system is much more than a cycle because it is affected by heat transfer and pressure drop (effect of thermophysical properties). Fig. 7 helps in preparing some answers. Thermodynamics does not look at the system and HXs so the asymptote for the cycle efficiency does not include heat transfer resistances in the HXs. Carnot cycle between evaporation and condensation temperatures equal to inlet air temperatures to evaporator and condenser provides the theoretically highest efficiency. Assuming infinite HX sizes as well as infinite heat source and sink and no pressure drop, isentropic compression along with dry suction and isenthalpic throttling the Rankine cycle based on air inlet temperatures (R_ai) defines the limit. Another idealization is the Rankine cycle based on air inlet temperature to the condenser (assuming infinite size) and air outlet from the evaporator (assuming the need for cold air or dehumidification) - R_aio. Reality is presented with COP_R_measured based on experimental data (still idealized for compressor efficiency). The arrows schematically present air temperature changes in the evaporator (light blue) and the condenser (pink). Fig. 8 presents the same comparison but on a T-s diagram for R134a at one condition (Taci = 40oC, Taei = 25oC) based on SAE AR CRP Phase 1 EnR134a system data. The actual cycle is in red.

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Figure 7. Relationship between theoretical limit (Carnot cycle – green square), Rankine cycle based on air inlet temperatures (R_ai), Rankine cycle based on air inlet to condenser and air outlet from evaporator temperatures

(R_aio), and measured values in experiment

Cond air inlet temp (40°C)

Evap air inlet temp (25°C)

Cond air outlet temp (51°C)

Evap air outlet temp (9°C)

Carnot cycle based on Teai and Tcai

Carnot cycle based on Tesat and Tcsat

Carnot cycle based on Teao and Tcai

Rankine cycle based on Tesat and Tcsat

Rankine cycle based on Teao and Tcai

Effective reservoir temp (46°C)


Carnot cycle based on effective reservoir temperatures

Actual cycle

Actual Cycle

Cond air inlet temp (40°C)

Evap air inlet temp (25°C)

Cond air outlet temp (51°C)

Evap air outlet temp (9°C)

Carnot cycle based on Teai and Tcai

Carnot cycle based on Tesat and Tcsat

Carnot cycle based on Teao and Tcai

Rankine cycle based on Tesat and Tcsat

Rankine cycle based on Teao and Tcai



Carnot cycle based on effective reservoir temperatures

Actual cycle

Actual Cycle

Figure 8. R134a analysis based on data from measurements in SAR AR CRP Phase 1 – E n R134a MAC system

T

s

Tcai

Teai

Tcro_sat

Teao

Tero_sat

COP_C_ai COP_R_aiCOP_R_aio

COP_R_measured

Theoretical limit

T

s

Tcai

Teai

Tcro_sat

Teao

Tero_sat

COP_C_ai COP_R_aiCOP_R_aio

COP_R_measured

Theoretical limit

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Table 3. SAE AR CRP Phase 1 Test matrix

The same COP analysis for numerous operating conditions presented in the SAE test matrix (Table 3) is conducted and the results are presented in Figure 9.

Comparison of the realized potential of the EnR134a system vs. the 2002 R744 MAC systems based on the Rankine Teao and Tcai for various operating conditions is presented in Fig. 10. It is obvious that CO2 achieved much more of its maximum in the current development. The logical question is: why, and would that be a limit for the future developments? A part of the answer is provided in the next section.

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EnR134a

ISH

70/5

ISH

60/5

ISH

45/5

LSH

45/5

HSH

45/5

ISH

60/E

ISH

45/E

LSH

45/E

HSH

45/E

ISH

50/5

IS

H35

/5LS

H35

/5H

SH35

/5H

SH35

/10

ISH

50/E

ISH

35/E

LSH

35/E

HS

H35

/EIS

H40

/5IS

H25

/5LS

H25

/5H

SH

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ISH

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0IS

H25

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LSH

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SH25

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ISH

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ISH

25/E

LSH

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ISH

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0IS

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/10_

80%

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COP_measured

COP_R_sat

COP_C_sat

COP_R_air_in_out

COP_C_air_in_out

COP_rev

COP_R_air_in

COP_C_air_in

ISH

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ISH

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HSH

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ISH

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/10_

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COP_rev

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COP_C_air_in

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/5H

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/5

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/10

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COP_C_air_in_out

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0

5

10

15

20

25

COP_measured COP_R_air_in_outCOP_R_sat COP_C_air_in_out

COP_C_air_in

R744

Figure 9. Comparison of COPs in various scenarios for En134a and 2002 R744 systems (test conditions in T. 3)

0.20

0.25

0.30

0.35

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0.45

0.50

ISH70/5

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/5

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/5

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/E

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/5

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/5

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/E

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LSL25/1

0

Pot

entia

l rea

lized

R134a

R744

Pot

entia

l rea

lized

R134a

R744

Figure 10. Realized potential of EnR134a and 2002 R744 MAC systems based on Rankine Teao and Tcai for various operating conditions provided in Table 3

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5 EVOLUTION OF SYSTEMS OVER THE YEARS AND THE REAL

OUTCOME Fig. 11 presents the effectiveness of the evaporators in several iterations of SAE AR CRP. The blue bars are the R134a evaporators and green bars are the R744 evaporators. The evaporators were exposed to the same operating conditions and were roughly the same size. Obviously the R744 evaporators were systematically more effective. The heat transfer coefficient on the refrigerant side is significantly higher for the CO2 but the analysis indicates that the air side is the dominant resistance so that is not the full explanation for the difference. Data indicate better distribution in R744 evaporators. That is probably due to different regrigerant flow control. The R744 system has a low-pressure receiver which provides slightly wet evaporator exit vapor covering dryout in DX systems and maldistribution effects. Refrigerant carried out from the accumulator evaporates in the IHX does not affect the refrigeration effect. The other reason, besides better refrigerant side heat transfer is less pressure drop and a more favorable temperature difference as schematically indicated in the Fig. 12.

Figure 13. Two microchannel gas coolers: cross flow and overall counter flow

Position along evaporator circuit

Loca

l tem

pera

ture

(t)

CO2

HFC

Air

t

Evaporating temp.to(psat,outlet)

Position along evaporator circuit

Loca

l tem

pera

ture

(t)

CO2

HFC

Air

t

Evaporating temp.to(psat,outlet)

0.00

0.10

0.20

0.30

0.40

0.50

0.60

0.70

0.80

0.90

1.00

I35a M 35a I25a M 25a I15 M 15

Test conditon

Effe

ctiv

enes

s [-

]

R134aR744

Fig. 11. Effectiveness of R134a and R744 evaporators at three stages of development

Fig. 12. Typical temperature profiles in R134a and R744 evaporators

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One of the major breakthroughs in R744 performance was when new design of overall counterflow gas cooler was implemented. Figure 13 shows two gas coolers that are different in circuiting only. One is crossflow (left and low) and the other (right and up) is a three slab overall counterflow. Results of both model and experiments show great advantages of counterflow arrangements Figure 14 presents experimental results of the first generation of the counterflow gas coolers. It is obvious that almost the entire heat exchange occurs in only the first third of the volume. Having in mind that the face area of the gas cooler was smaller than the baseline R134a condenser, and very low approaching temperature difference for a gas cooler of 0.5 – 1.0 ºC measured indicate further potentials for size reduction.

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Figure 14. Revealing temperature measurements in the counterflow gas cooler Not only heat exchanger effectiveness but also efficiencies of R744 compressors were constantly 5 – 10% points higher for R744 than R134a. Even in the IMAC 30/50 program when new, extremely efficient designs of compressors for R134a without suction valves (to reduce a major source of the inefficiencies) were explored isentropic efficiencies of R744 compressors were not met. 6. PROGRESS IN TIME Comparing the efficiencies of mobile air conditioning systems in Figures 5 and 15 one can see a dramatic increase in system efficiency over a decade of development in competition. It is interesting to notice that both R134a and R744 systems have been steadily improving. These improvements are much more significant internally, for systems made for each refrigerant, than the difference between two refrigerants at any point in time. That fact is presented in Figure 16. R744 systems were always lower in efficiency at higher ambient temperatures but higher at colder heat rejection conditions. The ambient conditions where COPs are equal are also indicated in Figure 15.

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Figure 16. Improvement of efficiency for each system over 10 years at any condition is greater than the difference in system efficiency at any given time

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7� CONCLUSIONS It is almost certain that R744 will not be implemented in the air conditioning systems in automobiles, at least not for some time. It is not because lower performance as expected at the beginning of the process but because of higher cost (estimates are not reliable but are in the range of 10%-30% higher compared to R134a systems, as a function of the source). The competition of two options: high-pressure transcritical R744 systems with low-pressure R134a and low-GWP candidates have dramatically improved the art and science of refrigeration. It was demonstrated that a very low performing R744 in Rankine cycle was dramatically improved by good designs that utilized favorable thermophysical properties of CO2 (higher specific heat, conductivity, vapor density, lower surface tension, etc.), and some smart cycle modification, mostly on the “left side”.

The question is how to benefit from the lessons learned in last decade and the half. Can we see another opportunity for CO2 or find the way to improve new synthetic refrigerants? One thing is certain: the most of the improvements in the R134a systems over last decade came from designs used first in CO2 systems:

microchannel evaporators were non existent for lower-pressure fluids but CO2 really needed them to maintain compactness at high working pressures,

internal heat exchangers (even working better for CO2 because of greater improvement potential for the cycle and lesser detrimental effects of pressure drop),

ejector systems are applied for R134a even though the effects are significantly lower, especially for expansion work recovery, etc.

That shows again that competition is the best way to empower creative thinking and develop or improve technology.

Stationary applications are less under pressure for global, unified refrigerant selection and almost unified solutions. Maybe that is one of the explanations why we see several successful applications for CO2 and other both natural and synthetic refrigerants.

We hope that proliferation of new ideas and creative solutions will follow the same trend as presented, enriching the profession, rejuvenating interest, and improving efficiency regardless of the refrigerant that will be chosen for any particular application.

REFERENCES [1] Pettersen, J. “Carbon Dioxide as a Refrigerant: Why is it Better than Expected?”, ASHRAE

meeting, Seminar 19, Transcritical CO2 Systems in Transportation Sector, June (1999) [2] Kim, M.H., J. Pettersen, C.W.Bullard, “Fundamental process and system design issues in CO2

vapor compression systems”. Progress in Energy and Combustion Science 30 119–174, (2004) [3] Yin, J., Y. Park, D. Boewe, R. McEnaney, A. Beaver, C. W. Bullard, and P. S. Hrnjak,

“Experimental and Model Comparison of Transcritical CO2 Versus R134a and R410 System Performance,” Proceedings of IIR Conference, Oslo, Norway, 376-385, June. (1998)

[4] Boewe, D. E., J. M. Yin, C. W. Bullard, and P. S. Hrnjak, “The Role of Suction Line Heat Exchanger in Transcritical R744 Mobile Air Conditioning Systems - Experimental and Model Analysis,” Proceedings of the SAE Conference, Detroit, MI, SAE Paper No. 1999-01-0583, March (1999).

[5] Elbel, S. W. and P. S. Hrnjak, “Experimental Validation of a Prototype Ejector Designed to Reduce Throttling Losses Encountered in Transcritical R744 System Operation,” International Journal of Refrigeration, 31:3, 411-422. (2008)

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