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Environmental implications of hydrocarbon refrigerants applied to the hermetic compressor
Materials & Design Materials and Design 26 (2005) 578–586 www.elsevier.com/locate/matdes
Environmental implications of hydrocarbon refrigerants applied to the hermetic compressor N.P. Garland *, M. Hadfield Sustainable Product Engineering Research Centre, Bournemouth University, UK Received 17 August 2004; accepted 18 August 2004
Abstract This paper describes the environmental impacts of hydrocarbon refrigerants deployed in the domestic refrigerator hermetic compressor. In-use durability is examined from a tribological viewpoint. Experimental tribological information is presented from physical test procedures involving sliding tests to establish wear mechanisms and friction coefficients within critical components. Hydrocarbon refrigerant R600a is compared with hydroflourocarbon R134a using aluminium on steel samples within a novel pressurised micro-friction test rig. The refrigerant R600a is tested for its influence upon the tribological performance of mineral oil (MO) and poly-ol-ester (POE) lubricant, whilst an R134a/POE charge combination is used as a benchmark. Tribological data is used to model long-term performance and subsequent environmental costs. 2004 Elsevier Ltd. All rights reserved. Keywords: Wear (E); Environmental performance (E); Abrasion (I)
1. Introduction Traditional refrigerants applied to domestic refrigeration applications have, until the mid 1990s, been chlorofluorocarbons (CFCs). The use of these compounds was a primary cause of ozone depletion in the upper atmosphere and were controlled, limited and finally phased out under the terms of the Montreal protocol [1]. CFCs used in domestic applications were initially replaced with hydrofluorocarbons (HFCs) which have a zero ozone depletion potential (ODP). The use of HFCs has, however, also come under scrutiny as they have a significant global warming potential (GWP) and have been replaced by hydrocarbon (HC) refrigerants in much of the domestic European and Asian markets [2]. The ODP of a refrigerant is an index number (by weight of product) with CFC refrigerant R11 rated at 1 (Table 1), whilst the *
Corresponding author. Tel.: +44 1202 524111. E-mail address: [email protected] (N.P. Garland).
0261-3069/$ - see front matter 2004 Elsevier Ltd. All rights reserved. doi:10.1016/j.matdes.2004.08.009
GWP relates to CO2 equivalence and all are classified under the Kyoto agreement [3] as greenhouse gases [4]. Whilst the refrigerants direct GWP contribution is only impacted by refrigerant leakage through failures in assembly, servicing or disposal procedures, the indirect contributions of a change in refrigerant can be much more significant. These include longevity through compressor wear and in-use energy consumption (CO2 emissions of power generation plants). The UK is currently obliged under the Kyoto agreement to reduce CO2 emissions by 121/2% compared to 1990 levels and has a domestic goal of 20% [5]. The use of hydrocarbon refrigerants has the direct environmental advantage of a greatly reduced GWP when compared to HFC refrigerants (Table 1). However the unseen or external environmental consequences are also relevant though somewhat harder to ascertain. Although HFCs have a much lower GWP and ODP compared to CFCs, the reduced coefficient of performance (COP) and inferior wear characteristics lead to a higher
N.P. Garland, M. Hadfield / Materials and Design 26 (2005) 578–586 Table 1 Refrigerant ozone depletion potential and global warming potential Refrigerant
Technology
ODP
GWP
R11 R12 R134a R717 (Ammonia) R600a (Isobutane) R290 (Propane) R600a/R290 Blend
CFC CFC HFC HC HC HC HC
1 1 0 0 0 0 0
4000 8500 1300 0 3 3 3
environmental burden during the products in use phase [6]. For HC R600a the COP is superior to both R134a and R12 [2,7–9], however the long term wear and durability of equipment using this refrigerant is unknown. For CFC refrigerants the lubricant of choice has been mineral oil (MO), for HFCs POE synthetic lubricants are preferred [10] while for HCs MOs are favoured, although POEs and polyalkylene glycol (PAG) are also compatible [7]. Lubrication within the compressor is not just a function of the lubricant, it is also governed by the selected refrigerant [11]. Wear and friction force within a MO lubricated contact reduces with the introduction of R12 refrigerant [10] but with POE and HFC in actual compressors the wear characteristics increase [11]. Experiments have shown that chlorine within the CFC refrigerant creates an extreme pressure (EP) film to reduce wear when boundary lubrication exists at the contact. With HFCs this same phenomenon occurs with the fluorine, only at higher loads. For the hermetic compressor, there are a number of areas where wear can occur. The main motor/crankshaft bearings and journals, the crank pin/connecting rod (big-end) bearing and journal, the piston pin/connecting (small-end) bearing and journal, the piston/cylinder sliding interface and the inlet/exhaust flapper (reed) valves are all subject to varying degrees of wear. Previous studies [10,12] have shown the piston pin/connecting rod bearing to be the most susceptible since, for the domestic refrigerator compressor, only the gudgeon pin/small end bearing operates within the boundary condition, the others being hydrodynamic journal bearings. The contact at the gudgeon pin/connecting rod is of the reciprocating conformal type, the surfaces effectively endure constant start/stop sliding motion and hence alternating lubrication regimes. Wear, occurring at the gudgeon pin/ small end bearing, will reduce the volume of refrigerant pumped per revolution, thereby increasing the duration of each operating cycle hence energy requirement. Friction at the same bearing, whilst not increasing the operating cycle, will absorb energy from the system.
2. Experimental micro-friction machine For the tribological evaluation of lubricant-refrigerant charge combinations under simulated opera-
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tional conditions a novel modified micro-friction machine was developed. The machine is based upon the existing Phoenix TE77 micro-friction machine but allows for control of the environmental conditions surrounding the contact to simulate those found within the hermetic compressor. The resultant TE57 micro-friction machine operates by sliding a lower plate sample in a reciprocating motion against a fixed sample pin or ball. The plate is mounted in a bath to hold the lubricant (if required) whilst the pin is connected via a feedback mechanism to a transducer to provide friction force feedback, hence friction coefficient (Fig. 1). The pin and plate samples can be electrically isolated so a potential divider may be used to measure voltage drop and hence lubricant film formation and contact gap. The samples and oil-bath are mounted within a pressure chamber to allow for controlled variation of the test environment. The chamber is further equipped with heating elements, a cooling system as well as charge pressure and temperature probes.
3. Experimental methodology 3.1. Overview For the pressurised micro-friction machine to provide meaningful results, the test regime should simulate the hermetic compressorÕs concentrated contact conditions as closely as possible. The contact pressure (load), temperature, operating pressure and surface velocities are selected to represent the conditions at the gudgeon pin/connecting rod interface. To ascertain the contact conditions present in the actual compressor a unit was dismantled and measurements taken or calculated as required. 3.2. Small end contact parameters The calculated contact condition indicates that the maximum Hertzian contact stress at the gudgeon pin is 55 MPa for the R134a compressor and 32.6 MPa for the R600a achieved at 160–180 after bottom dead centre (abdc). The small end to gudgeon pin surface speed of the contact however, is at a maximum across this range and at a minimum at 90 and 270. It is the reduction in surface speeds to zero during the reversal of sliding direction that cause the film thickness to reduce at these crank angles. The specific film thickness ratio (k) can be calculated from the film thickness and combined surface roughness indicating that boundary lubrication occurs (k<1) between 66 and 118 for the R134a compressor and 74 and 108 for the R600a.
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Fig. 1. System schematic.
Table 3 Operating environment
3.3. Test specimens Material test specimens were produced utilising materials and processes of a similar nature to those found in the actual compressor. In the case of the gudgeon pin, flat steel plate of an appropriate composition and hardness was selected (100/102Cr6) in surface finishes of 0.15 lRa (A type) and 0.05 lRa (B type). In the case of the connecting rod, an aluminium alloy was selected for the pin sample (LM13AE109). The cyclic loading of the actual contact complicates the task of matching the contact parameters. However, the load at the region of boundary lubrication is relatively constant with a pressure of 14–21 MPa from 66 to 118 for the R134a compressor and remaining constant at 11.5 MPa from 74 to 108 for the R600a. The load for the test specimens can therefore remain constant through each cycle (Table 2). 3.4. Operating environment The environmental conditions within the test chamber should match those found within the hermetic compressor at the gudgeon pin/small end contact (Table 3). From these tables a test program was determined for the evaluation of refrigerant lubricant combinations under a range of operational conditions (Table 4).
Table 2 Parameter selection
Equivalent radius (m) Crank angle () Load (N) Contact area · 106(m2) Contact stress (MPa) Surface velocity (m s1)
R134a
R600a
Test samples
0.27 70/110 35/63 2.5/3.3 14.2/19 0.098
0.39 70/110 43 3.7 11.6 0.12
0.005 70/110 20 0.28 70.3 0.13
Typical compressor operating temperature at start up Typical operating pressure at start up Typical compressor in-use temperature (discharge temp) Saturation pressure at 25 C (evaporator pressure) Compression pressure (sat pressure raised to comp temperature)
R600a
R134a
25 C
25 C
3.53 bar abs.
6.65 bar abs.
57 C
68 C
0.586 bar abs.
1.067 bar abs.
0.780 bar abs.
1.467 bar abs.
3.5. Test procedure The initial tests (tests 1–4) detailed in Table 4 were implemented using the pressurised micro-friction machine described previously. After installation of the test samples and pre-loading to 5 N, the pressure chamber was heated to the prescribed temperature. The measured quantity (2.5 ml) of lubricant was then added to the oil bath and the chamber sealed and evacuated prior to the introduction of the refrigerant atmosphere. With the charge conditions met, the machine was brought up to speed (27 Hz) and the load increased in 120 s, 5 N steps until the test load was reached and run for the specified duration before termination. Each of the tests being carried out a minimum of three times to provide average as well as raw data for subsequent analysis and interpretation. The initial round of testing was carried out using the 0.15 lm plates prior to supplanting with the 0.05 lm plates for extended duration testing (tests 5–12) and subsequent reduced load (15 N) testing (tests 13–16).
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Table 4 Experimental test conditions, + = unspecified lubricant additives Test no.
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16
Charge conditions
Experimental conditions
Temp (C)
Charge
57 57 57 68 57 57 57 57 57 57 68 68 57 57 68 68
R600a R600a R600a R134a R600a R600a R600a R600a R600a R600a R134a R134a R600a R600a R134a R134a
MO MO+ POE+ POE+ MO MO MO+ MO+ POE+ POE+ POE+ POE+ MO+ MO+ POE+ POE+
Pressure (bar)
Plate
Load (N)
Duration (s)
0.8 0.8 0.8 1.4 0.8 0.8 0.8 0.8 0.8 0.8 1.4 1.4 0.8 0.8 1.4 1.4
A A A A B B B B B B B B B B B B
20 20 20 20 20 20 20 20 20 20 20 20 15 15 15 15
7200 7200 7200 7200 14,400 86,400 14,400 86,400 14,400 86,400 14,400 86,400 86,400 237,600 86,400 237,600
4. Test results 4.1. Initial and extended experimental results The frictional results were recorded for each of the charge combinations and the average for each charge combination at each temperature range can be shown (Table 5 and Fig. 2). The wear scar dimensions at the pin were measured to provide further comparison between charge conditions and enabled calculation of the wear volume hence dimensional wear coefficient ‘‘k’’ (Eq. (1)) (Table 5 and Fig. 3). k ¼ V =ðLW Þ;
ð1Þ
where k is the dimensional wear coefficient (mm3(N m)1); V is the wear volume (mm3); L is the sliding distance (m); W is the load (N).
From Fig. 2 it can be observed that for the initial tests (tests 1–4) the plain mineral oil with R600a refrigerant charge out-performs all other combinations, although there is little more than a 10% difference between any charge combination with regard to friction. For the extended tests (tests 5–12) variance in frictional performance for each combination is again low with the standard deviation rising for the lower friction combinations. Again there is little variation in the results for the smooth plates over the shorter duration test range however, the POE charge combinations returned an improved frictional performance over the duration of the tests. From Fig. 3 it can be observed that for the initial tests (tests 1–4) the R134a additised POE combination generates up to 50 times as much wear volume as the base MO, hence higher wear coefficient. From the table it is
Table 5 Average and standard deviation of friction and wear coefficients Test no.
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16
Charge conditions
Friction coefficient
Duration
Charge
7200 7200 7200 7200 14,400 86,400 14,400 86,400 14,400 86,400 14,400 86,400 86,400 237,600 86,400 237,600
R600a R600a R600a R134a R600a R600a R600a R600a R600a R600a R134a R134a R600a R600a R134a R134a
MO MO+ POE+ POE+ MO MO MO+ MO+ POE+ POE+ POE+ POE+ MO MO POE+ POE+
Wear coefficient (k)
Mean
SD
Mean
SD
0.107 0.127 0.119 0.111 0.118 0.118 0.127 0.103 0.109 0.064 0.083 0.079 0.104 0.090 0.071 0.037
0.0039 0.0030 0.0047 0.0014 0.0007 0.0023 0.0066 0.0148 0.0100 0.0250 0.0139 0.0206 0.0054 0.0048 0.0139 0.0228
2.0E7 7.3E6 1.8E5 5.1E6 3.7E9 2.7E9 6.1E8 2.1E7 9.5E6 2.3E6 5.1E6 1.2E6 7.08E09 1.54E08 2.14E06 7.87E07
2.9E7 3.8E6 1.2E5 3.0E6 2.5E9 1.1E9 1.7E8 1.0E7 3.6E6 5.8E7 1.9E6 3.9E7 1.86E09 4.57E09 1.68E07 1.72E07
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friction coefficient
0.12 0.1 0.08 0.06 0.04 0.02
R600a M.O. R600a M.O. additised R600a P.O.E. additised R134a P.O.E. additised
0 A type plates 2hrs
B type plates 4hrs
B type plates 24hrs
test temperature Fig. 2. Average friction coefficient, initial and extended tests.
dimensional wear coefficient (k)
1.0E-04
1.0E-05
1.0E-06
1.0E-07
1.0E-08
R600a M.O. R600a M.O.add R600a P.O.E. add R134a P.O.E. add
1.0E-09 A type plates 2hrs
B type plates 4hrs
B type plates 24hrs
test condition Fig. 3. Average wear coefficient, initial and extended tests.
friction coefficient
0.14 0.12 0.1 0.08 0.06 0.04
R600a M.O. additised
0.02
R134a P.O.E. additised
0 A type plates 2hrs
B type plates 4hrs
B type plates 24hrs
B type plates 15N 24hrs
test temperature (˚C) Fig. 4. Average friction coefficient, reduced load tests.
B type plates 15N 66hrs
N.P. Garland, M. Hadfield / Materials and Design 26 (2005) 578–586
clear that whilst frictional performance can be predictable, the wear coefficient can vary considerably, the standard deviation being of a similar order to the average wear rate. For the extended duration tests (tests 5–12) the wear rate for both the additised and non-additised MO dropped by a factor of 100 whilst for POE charge combinations the reduction was less pronounced. The standard deviations for the wear coefficients are again high, indicating a wide range of wear performances. For the mineral oil based charge combinations the change in wear coefficient followed the change in plate surface finish. This would indicate that for these tests an abrasive wear regime is dominant. From the calculated wear coefficients it is clear that a change in surface roughness had little effect on the wear performance of the POE lubricant over these tests. 4.2. Reduced load test results
dimensional wear coefficient (k)
The reduced load test programme (tests 13–16), like the extended duration, focused on the compressorÕs typical operational conditions. For these tests the objective was to ascertain the wear regime of the POE charge combinations by extending the duration still further whilst reducing the contact pressure. As before, the friction and wear coefficients were averaged and plotted to charts (Figs. 4 and 5) for comparison with earlier results. Again, the reduction in friction coefficient for the MO charge combination is minimal but for the POE charge the results are much more dramatic. For the longer duration tests the friction drops off significantly over time, indicating the formation of a low friction regime at the contact. The standard deviation for the friction coefficient also increases for the POE charges as the coefficient falls, indicating that friction over these tests is less stable. The standard deviation for the wear rates is lower (relative to the wear coefficient) than those recorded for the previous tests,
indicating that wear over these tests is more stable. The change in the POE wear regime is born out by the chart of average wear coefficients. The wear chart differs from its predecessors by including a derived wear rate for the period 4–24 h in the case of 20 N extended and 24–66 h on the case of 15 N reduced load tests. From the chart it is clear that during these test periods the wear rate is greatly reduced for the additised POE charge combinations but remains relatively constant for the additised MO charge within each load range. In the case of the R600a additised MO, the wear coefficient rises to 2.0E8, whilst the R134a additised POE wear coefficient falls to 1.2E8. Microscope observations, completed using an Olympus BX60 device, of the extended test samples (Fig. 6) show material transfer from the softer pin to the plate sample and the generation of tribo-layers. This surface film is present on both the POE and the additised MO test plates but not the ones lubricated with lower wear rate non-additised MO. The worn pin surfaces of the non-additised MO lubricant tests appear to have also generated a boundary film unlike the other, much higher wear rate samples. Energy dispersive X-ray analysis, using a Jeol JXA840A electron probe micro-analyser, was carried out on pin and plate samples from test 25 (additised MO) and test 27 (additised POE). The results showed much higher levels of aluminium and silicon for the POE plates than for the MO, especially in the area surrounding the contact, where high levels of carbon were also detected. The spectra for the pin samples exhibited the presence of carbon and oxygen. The levels detected upon the MO pin were similar both within and surrounding the contact, whilst those of the POE were lower within the contact but higher surrounding. The higher aluminium and silicon presence at the POE steel plate indicates a greater level of transfer of material from the pin sample. Whilst the higher levels of carbon
1.0E-04
R600a M.O. add R134a P.O.E. add
1.0E-05
1.0E-06
1.0E-07
1.0E-08
1.0E-09 B plates 20N 4hrs
B plates 20N 24hrs
583
B plates 20N 4-24hrs
B plates 15N 24hrs
B plates 15N 66hrs
test condition Fig. 5. Average wear coefficient, reduced load tests.
B plates 15N 24-66hrs
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Fig. 6. Comparison of extended 4 h test pin (left) and plate (right): (a) R600a MO, (b) R600a MO+, (c) R600a POE+ and (d) R134a POE+.
are likely to be from a breakdown of the lubricant itself. The higher levels of carbon found at the MO pin contact indicate that any boundary layer developed is more durable than that developed by the POE. The presence of high levels of carbon, aluminium and silicon surrounding the POE plate contact may indicate the rapid removal of any boundary layer and subsequent deposition on the surrounding area. For the POE combinations, the dominant regime observed appears to be a corrosive wear mode, the aluminium pin sample reacting with the lubricant to generate a sacrificial film layer at the contact. At the beginning of the test the contact stress is high, therefore the film is easily removed by abrasive wear, hence high wear rate, and exposes a fresh surface for further reaction to take place. As the wear scar size increases, the contact stress reduces and the load is supported by the film, hence reduced friction and wear characteristics. The results are similar to those found for the friction and wear in the sliding of aluminium and steel pairs with P-containing additives [13], ZDPP [14] and chemical wear effects [15].
% duty cycle
100
5. Compressor life-cycle-analysis 5.1. Energy requirement The hermetic compressor operates over a typical duty cycle 30% whereby the device may run for 18 min for every hour of operation. Wear at the gudgeon pin/small end bearing will reduce the quantity of refrigerant pumped per revolution, thereby increasing the duty cycle over time (Fig. 7). The energy used by the compressor can be divided into useful power, used in the pumping operation of the compressor, and dissipated power, lost through friction and heat (Eq. (2)). P t ¼ P u þ P d;
where Pt = total energy (W); Pu = useful energy (W); Pd = dissipated energy (W). Useful energy used by the compressor can be calculated (Eq. (3)) by subtracting the energy recovered from the suction cycle from the energy required to compress and discharge the refrigerant.
R134a compressor R600a compressor
90 80 70 60 50 40 30 0
1
2
3
4
5
6
7
8
ð2Þ
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25
years of operation Fig. 7. Duty-cycles over time, R600a and R134a compressors.
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Table 6 LCA emissions for R600a and R134a type compressors,a manufacture,b recycling Key indicator
R134a compressor a
GWP (t/CO2 eq) SO2 (kg eq) NOx (kg eq) PM10 (kg eq)
R600a compressor
b
M and R
Use
Total
M and R
Use
Total
0.28 0.43 0.09 0.38
6.3 24.6 17.7 9.5
6.5 25 17.8 9.9
0.06 0.48 0.11 0.49
5.6 20.8 15.6 8.2
5.7 21.3 15.7 8.7
Table 7 Economic valuation of emissions for compressor 15 year LCA Key indicator
Economic valuation
Compressor type R134a
R600a
GWP SO2 NOx PM10
32.6 €/t 7.4 €/kg 2.3 €/kg 17 €/kg
213 185 41 169
186 158 36 148
Total
–
608
527
l P u ¼ ðF c F s Þ ; t
ð3Þ
where Fc = compression piston load (N); Fs = suction piston load (N); l = sliding distance (m); t = duration (s). Consumption of useful power will always be the same, since the duty cycle is extended to discharge the same volume of refrigerant. The dissipated power will, however, increase inline with the duty cycle, as the energy lost is relative to the operational duration. The total energy requirement over a 15-year lifecycle can therefore be calculated for both the R600a and R134a compressor. For the R600a compressor lubricated with additised MO the energy requirement is 4244 kW h and for the R134a additised POE compressor, 4541 kW h despite the lower friction and wear coefficients of the lubricant. 5.2. Environmental cost A lifecycle analysis (LCA) was carried out to ascertain the environmental consequences of each compressor type. The LCA encompassed the manufacture, use and disposal of the compressor over its 15-year life cycle
with measurements taken for key environmental indicators. For these works the key indicators, GWP (CO2), SO2, NOx, and PM10 are those used by the EU for energy sector emissions [16] (Table 6). Although UK electricity comes from a wide variety of sources, it can be viewed as a finite resource stock. The inventory may be continually renewed over time with older, less efficient stock being replaced by new, more efficient stock. The withdrawal of older stock may be delayed by increased demand, or advanced, by reductions in demand. For the inventory to be run with minimum emissions, low emitting stock such as renewables and nuclear would be run at full capacity with high emitting stock run as a top-up. Under these operating conditions the difference in energy requirements between the lower consumption R600a compressors and the higher consumption R134a compressors would be met entirely by higher emission stock. From the table it is clear that the manufacturing and recycling stages of the compressor LCA are much less significant than the use phase. Simply relating an avoided emission of 0.85 ton of CO2 to 3.7 kg of SO2 is meaningless until a common denomination is utilised therefore values have be attached to the quantity of pollutant emitted (Table 7). The values for SO2, NOx, and PM10 are taken from the Externe project [17] whilst the CO2 value has been taken from projected trading prices under the EU emissions trading scheme if the scheme is to be successful in achieving the requirements of the Kyoto Protocol [18]. Table 7 indicates that the value of emissions is €81 higher for the R134a compressor over the 15-year lifecycle of the device. More significantly, the R134a compressor performance deteriorates much quicker than the R600a (Table 8). The environmental burden more
Table 8 Economic valuation of annual emissions, extended life Key indicator
R134a
R600a
Yr 15
Yr 20
GWP SO2 NOx PM10
17 15 3.3 14
20 20 4.1 18
Total
49
62
Yr 25 32 35 7 30 105
Yr 15
Yr 20
Yr 25
13 11 2.6 11
14 12 2.8 11
16 13 3.1 13
38
41
45
586
N.P. Garland, M. Hadfield / Materials and Design 26 (2005) 578–586
than doubles if the R134a compressor is used beyond year 15 compared to a 20% increase for the R600a device.
6. Concluding remarks The performance of refrigerant/lubricant charge combinations are evaluated within the pressurised refrigerant environment. The best performing charge combinations, for wear, were those that established and sustained a boundary film upon the counter-face of the pin sample. The additised MO samples had a good boundary film on the pin and almost no wear at all on the plate, for these samples wear rates were low. In the case of the additised POE lubricant and refrigerant combinations deposition was present at the plate but not at the pin sample, under these conditions initial wear rates were high but reduced as the contact geometry became more favourable. The lubricant with the lowest wear rate at the pin counter-face was the non-additised MO. The higher contact stress and reduced gudgeon pin bearing area of the R134a compared with the R600a compressor leads to reduced performance over time, despite the lower wear rate achieved for the R134a POE charge. The R600a compressor continues to perform at the end of its 15-year cycle where as the performance of the R134a compressor deteriorates rapidly at, and beyond, 15 years. This fall off in performance results in a greater environmental burden compared to the R600a.
Acknowledgement The authors acknowledge the Engineering and Physical Science Research Council (EPSRC), UK for funding this research project as part of the Design for Whole Lifecycle Programme.
References [1] UNEP, Montreal Protocol on Substances that Deplete the Ozone Layer. New York: United Nations; 1987. [2] Meyer A. Hydrocarbon blends in small scale refrigeration appliances. GTZ Yearbook 1996 1996:87–92. [3] UNFCC, Kyoto Protocol to the United Nations Frame Convention of Climate Change. New York: United Nations; 1997. [4] Cranvey D. Replacement refrigerants – environment based fluid selection. Mech. Incorporated Eng. 1997:129–33. [5] Royal Commission on Environmental Pollution, Energy – The Changing Climate. London: Royal Commission; 2001. [6] Ciantar C, Hadfield M. An environmental evaluation of mechanical systems using environmentally acceptable refrigerants. Int. J. Life Cycle Assess. 2000;5(4):209–20. [7] Granryd E. Hydrocarbons as refrigerants. Int. J. Refrigerat. 2000;24:15–24. [8] Meurer C, Law J. R12 replacements for the appliance industry. In: China Refrigeration 1998, Solkane Refrigerant Seminar, Shanghai, China, 1998. [9] Paul J. A fresh look at hydrocarbon refrigeration – experience and outlook. GTZ Yearbook 1996 1996:73–84. [10] Na BC, Chun KJ, Han D-C. A tribological study of refrigeration oils under HFC-134a environment. Tribol. Int. 1998;30(9):707–16. [11] Mizuhara K, Akei M, Matsuzaki T. The friction and wear behaviour in controlled alternative refrigerant atmosphere. STLE Tribol. Trans. 1994;37(1):120–8. [12] Ciantar C, Hadfield M, Smith AM, Swallow A. The Influences of lubricant viscosity on the wear of hermetic compressor components in HFC-134a environments. Wear 1999;236:1–8. [13] Yong Wan, Qunji X. Friction and wear characteristics of Pcontaining antiwear and extreme pressure additives in the sliding of steel against aluminium alloy. Wear 1995;188(1/2):27–32. [14] Yong W, Qunji X, Lile C. Friction and wear characteristics of ZDDP in the sliding of steel against aluminium alloy. Tribol. Int. 1997;30(10):767–72. [15] Montgomery RS. Chemical effects on wear in the lubrication of aluminium. Wear 1965;8(4):289–302. [16] European Commission, ExternE- Externalities of Energy, Methodology Annexes. 1998, European Commission Directorate General XII Science, Research and Development. [17] European Commission, ExternE- Externalities of Energy, UK Introduction. 1998, Department for Environment, Food and Rural Affairs. [18] Capros P, Mantzos L. The economic effects of EU-wide industrylevel emissions trading to reduce greenhouse gases. Athens: Institute of Communication and Computer Systems of National Technical University of Athens; 2000.
Environmental implications of hydrocarbon refrigerants applied to the hermetic compressor N.P. Garland *, M. Hadfield Sustainable Product Engineering Research Centre, Bournemouth University, UK Received 17 August 2004; accepted 18 August 2004
Abstract This paper describes the environmental impacts of hydrocarbon refrigerants deployed in the domestic refrigerator hermetic compressor. In-use durability is examined from a tribological viewpoint. Experimental tribological information is presented from physical test procedures involving sliding tests to establish wear mechanisms and friction coefficients within critical components. Hydrocarbon refrigerant R600a is compared with hydroflourocarbon R134a using aluminium on steel samples within a novel pressurised micro-friction test rig. The refrigerant R600a is tested for its influence upon the tribological performance of mineral oil (MO) and poly-ol-ester (POE) lubricant, whilst an R134a/POE charge combination is used as a benchmark. Tribological data is used to model long-term performance and subsequent environmental costs. 2004 Elsevier Ltd. All rights reserved. Keywords: Wear (E); Environmental performance (E); Abrasion (I)
1. Introduction Traditional refrigerants applied to domestic refrigeration applications have, until the mid 1990s, been chlorofluorocarbons (CFCs). The use of these compounds was a primary cause of ozone depletion in the upper atmosphere and were controlled, limited and finally phased out under the terms of the Montreal protocol [1]. CFCs used in domestic applications were initially replaced with hydrofluorocarbons (HFCs) which have a zero ozone depletion potential (ODP). The use of HFCs has, however, also come under scrutiny as they have a significant global warming potential (GWP) and have been replaced by hydrocarbon (HC) refrigerants in much of the domestic European and Asian markets [2]. The ODP of a refrigerant is an index number (by weight of product) with CFC refrigerant R11 rated at 1 (Table 1), whilst the *
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0261-3069/$ - see front matter 2004 Elsevier Ltd. All rights reserved. doi:10.1016/j.matdes.2004.08.009
GWP relates to CO2 equivalence and all are classified under the Kyoto agreement [3] as greenhouse gases [4]. Whilst the refrigerants direct GWP contribution is only impacted by refrigerant leakage through failures in assembly, servicing or disposal procedures, the indirect contributions of a change in refrigerant can be much more significant. These include longevity through compressor wear and in-use energy consumption (CO2 emissions of power generation plants). The UK is currently obliged under the Kyoto agreement to reduce CO2 emissions by 121/2% compared to 1990 levels and has a domestic goal of 20% [5]. The use of hydrocarbon refrigerants has the direct environmental advantage of a greatly reduced GWP when compared to HFC refrigerants (Table 1). However the unseen or external environmental consequences are also relevant though somewhat harder to ascertain. Although HFCs have a much lower GWP and ODP compared to CFCs, the reduced coefficient of performance (COP) and inferior wear characteristics lead to a higher
N.P. Garland, M. Hadfield / Materials and Design 26 (2005) 578–586 Table 1 Refrigerant ozone depletion potential and global warming potential Refrigerant
Technology
ODP
GWP
R11 R12 R134a R717 (Ammonia) R600a (Isobutane) R290 (Propane) R600a/R290 Blend
CFC CFC HFC HC HC HC HC
1 1 0 0 0 0 0
4000 8500 1300 0 3 3 3
environmental burden during the products in use phase [6]. For HC R600a the COP is superior to both R134a and R12 [2,7–9], however the long term wear and durability of equipment using this refrigerant is unknown. For CFC refrigerants the lubricant of choice has been mineral oil (MO), for HFCs POE synthetic lubricants are preferred [10] while for HCs MOs are favoured, although POEs and polyalkylene glycol (PAG) are also compatible [7]. Lubrication within the compressor is not just a function of the lubricant, it is also governed by the selected refrigerant [11]. Wear and friction force within a MO lubricated contact reduces with the introduction of R12 refrigerant [10] but with POE and HFC in actual compressors the wear characteristics increase [11]. Experiments have shown that chlorine within the CFC refrigerant creates an extreme pressure (EP) film to reduce wear when boundary lubrication exists at the contact. With HFCs this same phenomenon occurs with the fluorine, only at higher loads. For the hermetic compressor, there are a number of areas where wear can occur. The main motor/crankshaft bearings and journals, the crank pin/connecting rod (big-end) bearing and journal, the piston pin/connecting (small-end) bearing and journal, the piston/cylinder sliding interface and the inlet/exhaust flapper (reed) valves are all subject to varying degrees of wear. Previous studies [10,12] have shown the piston pin/connecting rod bearing to be the most susceptible since, for the domestic refrigerator compressor, only the gudgeon pin/small end bearing operates within the boundary condition, the others being hydrodynamic journal bearings. The contact at the gudgeon pin/connecting rod is of the reciprocating conformal type, the surfaces effectively endure constant start/stop sliding motion and hence alternating lubrication regimes. Wear, occurring at the gudgeon pin/ small end bearing, will reduce the volume of refrigerant pumped per revolution, thereby increasing the duration of each operating cycle hence energy requirement. Friction at the same bearing, whilst not increasing the operating cycle, will absorb energy from the system.
2. Experimental micro-friction machine For the tribological evaluation of lubricant-refrigerant charge combinations under simulated opera-
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tional conditions a novel modified micro-friction machine was developed. The machine is based upon the existing Phoenix TE77 micro-friction machine but allows for control of the environmental conditions surrounding the contact to simulate those found within the hermetic compressor. The resultant TE57 micro-friction machine operates by sliding a lower plate sample in a reciprocating motion against a fixed sample pin or ball. The plate is mounted in a bath to hold the lubricant (if required) whilst the pin is connected via a feedback mechanism to a transducer to provide friction force feedback, hence friction coefficient (Fig. 1). The pin and plate samples can be electrically isolated so a potential divider may be used to measure voltage drop and hence lubricant film formation and contact gap. The samples and oil-bath are mounted within a pressure chamber to allow for controlled variation of the test environment. The chamber is further equipped with heating elements, a cooling system as well as charge pressure and temperature probes.
3. Experimental methodology 3.1. Overview For the pressurised micro-friction machine to provide meaningful results, the test regime should simulate the hermetic compressorÕs concentrated contact conditions as closely as possible. The contact pressure (load), temperature, operating pressure and surface velocities are selected to represent the conditions at the gudgeon pin/connecting rod interface. To ascertain the contact conditions present in the actual compressor a unit was dismantled and measurements taken or calculated as required. 3.2. Small end contact parameters The calculated contact condition indicates that the maximum Hertzian contact stress at the gudgeon pin is 55 MPa for the R134a compressor and 32.6 MPa for the R600a achieved at 160–180 after bottom dead centre (abdc). The small end to gudgeon pin surface speed of the contact however, is at a maximum across this range and at a minimum at 90 and 270. It is the reduction in surface speeds to zero during the reversal of sliding direction that cause the film thickness to reduce at these crank angles. The specific film thickness ratio (k) can be calculated from the film thickness and combined surface roughness indicating that boundary lubrication occurs (k<1) between 66 and 118 for the R134a compressor and 74 and 108 for the R600a.
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Fig. 1. System schematic.
Table 3 Operating environment
3.3. Test specimens Material test specimens were produced utilising materials and processes of a similar nature to those found in the actual compressor. In the case of the gudgeon pin, flat steel plate of an appropriate composition and hardness was selected (100/102Cr6) in surface finishes of 0.15 lRa (A type) and 0.05 lRa (B type). In the case of the connecting rod, an aluminium alloy was selected for the pin sample (LM13AE109). The cyclic loading of the actual contact complicates the task of matching the contact parameters. However, the load at the region of boundary lubrication is relatively constant with a pressure of 14–21 MPa from 66 to 118 for the R134a compressor and remaining constant at 11.5 MPa from 74 to 108 for the R600a. The load for the test specimens can therefore remain constant through each cycle (Table 2). 3.4. Operating environment The environmental conditions within the test chamber should match those found within the hermetic compressor at the gudgeon pin/small end contact (Table 3). From these tables a test program was determined for the evaluation of refrigerant lubricant combinations under a range of operational conditions (Table 4).
Table 2 Parameter selection
Equivalent radius (m) Crank angle () Load (N) Contact area · 106(m2) Contact stress (MPa) Surface velocity (m s1)
R134a
R600a
Test samples
0.27 70/110 35/63 2.5/3.3 14.2/19 0.098
0.39 70/110 43 3.7 11.6 0.12
0.005 70/110 20 0.28 70.3 0.13
Typical compressor operating temperature at start up Typical operating pressure at start up Typical compressor in-use temperature (discharge temp) Saturation pressure at 25 C (evaporator pressure) Compression pressure (sat pressure raised to comp temperature)
R600a
R134a
25 C
25 C
3.53 bar abs.
6.65 bar abs.
57 C
68 C
0.586 bar abs.
1.067 bar abs.
0.780 bar abs.
1.467 bar abs.
3.5. Test procedure The initial tests (tests 1–4) detailed in Table 4 were implemented using the pressurised micro-friction machine described previously. After installation of the test samples and pre-loading to 5 N, the pressure chamber was heated to the prescribed temperature. The measured quantity (2.5 ml) of lubricant was then added to the oil bath and the chamber sealed and evacuated prior to the introduction of the refrigerant atmosphere. With the charge conditions met, the machine was brought up to speed (27 Hz) and the load increased in 120 s, 5 N steps until the test load was reached and run for the specified duration before termination. Each of the tests being carried out a minimum of three times to provide average as well as raw data for subsequent analysis and interpretation. The initial round of testing was carried out using the 0.15 lm plates prior to supplanting with the 0.05 lm plates for extended duration testing (tests 5–12) and subsequent reduced load (15 N) testing (tests 13–16).
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Table 4 Experimental test conditions, + = unspecified lubricant additives Test no.
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16
Charge conditions
Experimental conditions
Temp (C)
Charge
57 57 57 68 57 57 57 57 57 57 68 68 57 57 68 68
R600a R600a R600a R134a R600a R600a R600a R600a R600a R600a R134a R134a R600a R600a R134a R134a
MO MO+ POE+ POE+ MO MO MO+ MO+ POE+ POE+ POE+ POE+ MO+ MO+ POE+ POE+
Pressure (bar)
Plate
Load (N)
Duration (s)
0.8 0.8 0.8 1.4 0.8 0.8 0.8 0.8 0.8 0.8 1.4 1.4 0.8 0.8 1.4 1.4
A A A A B B B B B B B B B B B B
20 20 20 20 20 20 20 20 20 20 20 20 15 15 15 15
7200 7200 7200 7200 14,400 86,400 14,400 86,400 14,400 86,400 14,400 86,400 86,400 237,600 86,400 237,600
4. Test results 4.1. Initial and extended experimental results The frictional results were recorded for each of the charge combinations and the average for each charge combination at each temperature range can be shown (Table 5 and Fig. 2). The wear scar dimensions at the pin were measured to provide further comparison between charge conditions and enabled calculation of the wear volume hence dimensional wear coefficient ‘‘k’’ (Eq. (1)) (Table 5 and Fig. 3). k ¼ V =ðLW Þ;
ð1Þ
where k is the dimensional wear coefficient (mm3(N m)1); V is the wear volume (mm3); L is the sliding distance (m); W is the load (N).
From Fig. 2 it can be observed that for the initial tests (tests 1–4) the plain mineral oil with R600a refrigerant charge out-performs all other combinations, although there is little more than a 10% difference between any charge combination with regard to friction. For the extended tests (tests 5–12) variance in frictional performance for each combination is again low with the standard deviation rising for the lower friction combinations. Again there is little variation in the results for the smooth plates over the shorter duration test range however, the POE charge combinations returned an improved frictional performance over the duration of the tests. From Fig. 3 it can be observed that for the initial tests (tests 1–4) the R134a additised POE combination generates up to 50 times as much wear volume as the base MO, hence higher wear coefficient. From the table it is
Table 5 Average and standard deviation of friction and wear coefficients Test no.
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16
Charge conditions
Friction coefficient
Duration
Charge
7200 7200 7200 7200 14,400 86,400 14,400 86,400 14,400 86,400 14,400 86,400 86,400 237,600 86,400 237,600
R600a R600a R600a R134a R600a R600a R600a R600a R600a R600a R134a R134a R600a R600a R134a R134a
MO MO+ POE+ POE+ MO MO MO+ MO+ POE+ POE+ POE+ POE+ MO MO POE+ POE+
Wear coefficient (k)
Mean
SD
Mean
SD
0.107 0.127 0.119 0.111 0.118 0.118 0.127 0.103 0.109 0.064 0.083 0.079 0.104 0.090 0.071 0.037
0.0039 0.0030 0.0047 0.0014 0.0007 0.0023 0.0066 0.0148 0.0100 0.0250 0.0139 0.0206 0.0054 0.0048 0.0139 0.0228
2.0E7 7.3E6 1.8E5 5.1E6 3.7E9 2.7E9 6.1E8 2.1E7 9.5E6 2.3E6 5.1E6 1.2E6 7.08E09 1.54E08 2.14E06 7.87E07
2.9E7 3.8E6 1.2E5 3.0E6 2.5E9 1.1E9 1.7E8 1.0E7 3.6E6 5.8E7 1.9E6 3.9E7 1.86E09 4.57E09 1.68E07 1.72E07
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friction coefficient
0.12 0.1 0.08 0.06 0.04 0.02
R600a M.O. R600a M.O. additised R600a P.O.E. additised R134a P.O.E. additised
0 A type plates 2hrs
B type plates 4hrs
B type plates 24hrs
test temperature Fig. 2. Average friction coefficient, initial and extended tests.
dimensional wear coefficient (k)
1.0E-04
1.0E-05
1.0E-06
1.0E-07
1.0E-08
R600a M.O. R600a M.O.add R600a P.O.E. add R134a P.O.E. add
1.0E-09 A type plates 2hrs
B type plates 4hrs
B type plates 24hrs
test condition Fig. 3. Average wear coefficient, initial and extended tests.
friction coefficient
0.14 0.12 0.1 0.08 0.06 0.04
R600a M.O. additised
0.02
R134a P.O.E. additised
0 A type plates 2hrs
B type plates 4hrs
B type plates 24hrs
B type plates 15N 24hrs
test temperature (˚C) Fig. 4. Average friction coefficient, reduced load tests.
B type plates 15N 66hrs
N.P. Garland, M. Hadfield / Materials and Design 26 (2005) 578–586
clear that whilst frictional performance can be predictable, the wear coefficient can vary considerably, the standard deviation being of a similar order to the average wear rate. For the extended duration tests (tests 5–12) the wear rate for both the additised and non-additised MO dropped by a factor of 100 whilst for POE charge combinations the reduction was less pronounced. The standard deviations for the wear coefficients are again high, indicating a wide range of wear performances. For the mineral oil based charge combinations the change in wear coefficient followed the change in plate surface finish. This would indicate that for these tests an abrasive wear regime is dominant. From the calculated wear coefficients it is clear that a change in surface roughness had little effect on the wear performance of the POE lubricant over these tests. 4.2. Reduced load test results
dimensional wear coefficient (k)
The reduced load test programme (tests 13–16), like the extended duration, focused on the compressorÕs typical operational conditions. For these tests the objective was to ascertain the wear regime of the POE charge combinations by extending the duration still further whilst reducing the contact pressure. As before, the friction and wear coefficients were averaged and plotted to charts (Figs. 4 and 5) for comparison with earlier results. Again, the reduction in friction coefficient for the MO charge combination is minimal but for the POE charge the results are much more dramatic. For the longer duration tests the friction drops off significantly over time, indicating the formation of a low friction regime at the contact. The standard deviation for the friction coefficient also increases for the POE charges as the coefficient falls, indicating that friction over these tests is less stable. The standard deviation for the wear rates is lower (relative to the wear coefficient) than those recorded for the previous tests,
indicating that wear over these tests is more stable. The change in the POE wear regime is born out by the chart of average wear coefficients. The wear chart differs from its predecessors by including a derived wear rate for the period 4–24 h in the case of 20 N extended and 24–66 h on the case of 15 N reduced load tests. From the chart it is clear that during these test periods the wear rate is greatly reduced for the additised POE charge combinations but remains relatively constant for the additised MO charge within each load range. In the case of the R600a additised MO, the wear coefficient rises to 2.0E8, whilst the R134a additised POE wear coefficient falls to 1.2E8. Microscope observations, completed using an Olympus BX60 device, of the extended test samples (Fig. 6) show material transfer from the softer pin to the plate sample and the generation of tribo-layers. This surface film is present on both the POE and the additised MO test plates but not the ones lubricated with lower wear rate non-additised MO. The worn pin surfaces of the non-additised MO lubricant tests appear to have also generated a boundary film unlike the other, much higher wear rate samples. Energy dispersive X-ray analysis, using a Jeol JXA840A electron probe micro-analyser, was carried out on pin and plate samples from test 25 (additised MO) and test 27 (additised POE). The results showed much higher levels of aluminium and silicon for the POE plates than for the MO, especially in the area surrounding the contact, where high levels of carbon were also detected. The spectra for the pin samples exhibited the presence of carbon and oxygen. The levels detected upon the MO pin were similar both within and surrounding the contact, whilst those of the POE were lower within the contact but higher surrounding. The higher aluminium and silicon presence at the POE steel plate indicates a greater level of transfer of material from the pin sample. Whilst the higher levels of carbon
1.0E-04
R600a M.O. add R134a P.O.E. add
1.0E-05
1.0E-06
1.0E-07
1.0E-08
1.0E-09 B plates 20N 4hrs
B plates 20N 24hrs
583
B plates 20N 4-24hrs
B plates 15N 24hrs
B plates 15N 66hrs
test condition Fig. 5. Average wear coefficient, reduced load tests.
B plates 15N 24-66hrs
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Fig. 6. Comparison of extended 4 h test pin (left) and plate (right): (a) R600a MO, (b) R600a MO+, (c) R600a POE+ and (d) R134a POE+.
are likely to be from a breakdown of the lubricant itself. The higher levels of carbon found at the MO pin contact indicate that any boundary layer developed is more durable than that developed by the POE. The presence of high levels of carbon, aluminium and silicon surrounding the POE plate contact may indicate the rapid removal of any boundary layer and subsequent deposition on the surrounding area. For the POE combinations, the dominant regime observed appears to be a corrosive wear mode, the aluminium pin sample reacting with the lubricant to generate a sacrificial film layer at the contact. At the beginning of the test the contact stress is high, therefore the film is easily removed by abrasive wear, hence high wear rate, and exposes a fresh surface for further reaction to take place. As the wear scar size increases, the contact stress reduces and the load is supported by the film, hence reduced friction and wear characteristics. The results are similar to those found for the friction and wear in the sliding of aluminium and steel pairs with P-containing additives [13], ZDPP [14] and chemical wear effects [15].
% duty cycle
100
5. Compressor life-cycle-analysis 5.1. Energy requirement The hermetic compressor operates over a typical duty cycle 30% whereby the device may run for 18 min for every hour of operation. Wear at the gudgeon pin/small end bearing will reduce the quantity of refrigerant pumped per revolution, thereby increasing the duty cycle over time (Fig. 7). The energy used by the compressor can be divided into useful power, used in the pumping operation of the compressor, and dissipated power, lost through friction and heat (Eq. (2)). P t ¼ P u þ P d;
where Pt = total energy (W); Pu = useful energy (W); Pd = dissipated energy (W). Useful energy used by the compressor can be calculated (Eq. (3)) by subtracting the energy recovered from the suction cycle from the energy required to compress and discharge the refrigerant.
R134a compressor R600a compressor
90 80 70 60 50 40 30 0
1
2
3
4
5
6
7
8
ð2Þ
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25
years of operation Fig. 7. Duty-cycles over time, R600a and R134a compressors.
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Table 6 LCA emissions for R600a and R134a type compressors,a manufacture,b recycling Key indicator
R134a compressor a
GWP (t/CO2 eq) SO2 (kg eq) NOx (kg eq) PM10 (kg eq)
R600a compressor
b
M and R
Use
Total
M and R
Use
Total
0.28 0.43 0.09 0.38
6.3 24.6 17.7 9.5
6.5 25 17.8 9.9
0.06 0.48 0.11 0.49
5.6 20.8 15.6 8.2
5.7 21.3 15.7 8.7
Table 7 Economic valuation of emissions for compressor 15 year LCA Key indicator
Economic valuation
Compressor type R134a
R600a
GWP SO2 NOx PM10
32.6 €/t 7.4 €/kg 2.3 €/kg 17 €/kg
213 185 41 169
186 158 36 148
Total
–
608
527
l P u ¼ ðF c F s Þ ; t
ð3Þ
where Fc = compression piston load (N); Fs = suction piston load (N); l = sliding distance (m); t = duration (s). Consumption of useful power will always be the same, since the duty cycle is extended to discharge the same volume of refrigerant. The dissipated power will, however, increase inline with the duty cycle, as the energy lost is relative to the operational duration. The total energy requirement over a 15-year lifecycle can therefore be calculated for both the R600a and R134a compressor. For the R600a compressor lubricated with additised MO the energy requirement is 4244 kW h and for the R134a additised POE compressor, 4541 kW h despite the lower friction and wear coefficients of the lubricant. 5.2. Environmental cost A lifecycle analysis (LCA) was carried out to ascertain the environmental consequences of each compressor type. The LCA encompassed the manufacture, use and disposal of the compressor over its 15-year life cycle
with measurements taken for key environmental indicators. For these works the key indicators, GWP (CO2), SO2, NOx, and PM10 are those used by the EU for energy sector emissions [16] (Table 6). Although UK electricity comes from a wide variety of sources, it can be viewed as a finite resource stock. The inventory may be continually renewed over time with older, less efficient stock being replaced by new, more efficient stock. The withdrawal of older stock may be delayed by increased demand, or advanced, by reductions in demand. For the inventory to be run with minimum emissions, low emitting stock such as renewables and nuclear would be run at full capacity with high emitting stock run as a top-up. Under these operating conditions the difference in energy requirements between the lower consumption R600a compressors and the higher consumption R134a compressors would be met entirely by higher emission stock. From the table it is clear that the manufacturing and recycling stages of the compressor LCA are much less significant than the use phase. Simply relating an avoided emission of 0.85 ton of CO2 to 3.7 kg of SO2 is meaningless until a common denomination is utilised therefore values have be attached to the quantity of pollutant emitted (Table 7). The values for SO2, NOx, and PM10 are taken from the Externe project [17] whilst the CO2 value has been taken from projected trading prices under the EU emissions trading scheme if the scheme is to be successful in achieving the requirements of the Kyoto Protocol [18]. Table 7 indicates that the value of emissions is €81 higher for the R134a compressor over the 15-year lifecycle of the device. More significantly, the R134a compressor performance deteriorates much quicker than the R600a (Table 8). The environmental burden more
Table 8 Economic valuation of annual emissions, extended life Key indicator
R134a
R600a
Yr 15
Yr 20
GWP SO2 NOx PM10
17 15 3.3 14
20 20 4.1 18
Total
49
62
Yr 25 32 35 7 30 105
Yr 15
Yr 20
Yr 25
13 11 2.6 11
14 12 2.8 11
16 13 3.1 13
38
41
45
586
N.P. Garland, M. Hadfield / Materials and Design 26 (2005) 578–586
than doubles if the R134a compressor is used beyond year 15 compared to a 20% increase for the R600a device.
6. Concluding remarks The performance of refrigerant/lubricant charge combinations are evaluated within the pressurised refrigerant environment. The best performing charge combinations, for wear, were those that established and sustained a boundary film upon the counter-face of the pin sample. The additised MO samples had a good boundary film on the pin and almost no wear at all on the plate, for these samples wear rates were low. In the case of the additised POE lubricant and refrigerant combinations deposition was present at the plate but not at the pin sample, under these conditions initial wear rates were high but reduced as the contact geometry became more favourable. The lubricant with the lowest wear rate at the pin counter-face was the non-additised MO. The higher contact stress and reduced gudgeon pin bearing area of the R134a compared with the R600a compressor leads to reduced performance over time, despite the lower wear rate achieved for the R134a POE charge. The R600a compressor continues to perform at the end of its 15-year cycle where as the performance of the R134a compressor deteriorates rapidly at, and beyond, 15 years. This fall off in performance results in a greater environmental burden compared to the R600a.
Acknowledgement The authors acknowledge the Engineering and Physical Science Research Council (EPSRC), UK for funding this research project as part of the Design for Whole Lifecycle Programme.
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