United States Environmental Protection Agency Office of Atmospheric and Indoor Air Programs EPA-430-R-93-010 June 1993

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### State OfThe Art Survey Of Hermetic CompressorTechnology Applicable to Domestic Refrigerator/Freezers

Prepared for Environmental Protection Agency Division of Global Change

Revised: March, 1993

Reference 64128 64122 67986Prologue

Compressortechnology will play an importantrole in allowing potential efficiency improvements ofvarious refrigerator/freezer design alternatives. This review ofthe technology ofsmall hermetic refrigeration compressors was undertaken as part ofa larger study to evaluate the options for maximizing the efficiency ofdomestic refrigerators. The compressorcharacteristics documented in the report are being used in the overall study as partofthe inputdatabase for modeling and evaluating design options. Findings 1. The efficiencyof "large" compressors (capacitygreaterthan 750 Btu/hr) has been improvedsignificantlyoverthepastdecade. Furtherimprovements are possible, atincrementallyhighercosts.

• In 1980 the best "large" compressor energy efficiency ratio (EER), at standard conditions, was 4.0 BtuIWatt-hr.

• The bestcurrentproduction EERlevels are 5.5.

• In response to 1993 refrigerator/freezerefficiency standards, compressor models with significantly upgraded efficiency are becoming available (for use with CFC-12).

• Compressors with nominal capacity above 750 Btu/hr, with EER levels up to 6.0 will be commercially available to OEM's in 1993.

• This represents approximately a 6% increase in the efficiency level ofcommercially available compressors in this capacity range. About one halfofthis increase is attributable to higher motor efficiencies, the remainder to incremental reductions in mechanical and thermal losses.

• An EER level ofapproximately 6.5 is technically feasible, at an incremental increase in OEM costs on the order of$15. 2. The value ofthese energysavings was calculated.

• Assuming an electric energy cost of8¢ per KwH, for real discount rates less than 10%, the real, present value ofthe saved energy is on the order of$100, over the projected 15 year life ofthe refrigerator, much greater than the incremental cost, of the increased efficiency compressor. 3. Smallercompressors will be neededin future refrigerators.

• Superinsulation (either vacuum panels or thicker walls) is likely to reduce loads." Dual loop (or staged) systems using two compressors may replace single compressor/evaporatordesigns to gain theoretical energy efficiency advantages. 4. Smallcompressors are inefficientbecause of

### a lackofeconomic incentive 

### to develop efficientmodels; substantialimprovements can be obtainedby applying currentlyavailable technology.

" Current small compressors are relatively inefficient, mainly due to inefficientmotors and high levels ofsuction gas heating.

" Currenttechnology exists to make efficient motors for small compressors at a reasonable price.

" It is technically feasible to develop and produce an efficient small compressor, having an EER greater than 5.0 5. Energyimprovements from improvedsmallcompressorperformance are possible with improvedmotors andapplication ofotherexisting technology. Dualloop andsuper-insulation systems shouldbe economically viable with this currentlyavailable (butas ofyet) unmanufactured technology.

• Small compressors can beimproved to efficiency levels within 10 to 15 percent of large compressors for approximately $15.00; the energy gain they would produce in a dual loop system would be worth roughly $100.

• Research and improved manufacturing could improve the benefit/costcalculations used in this report.

iiTable of Contents 1.0 Introduction 1-1 1.1 Stratospheric Ozone Depletion and the CFC Phase Out 1-1 1.2 Global Warming 1-2 1.3 DOE Appliance Energy Efficiency Standard Setting 1-2 1.4 OtherRegulatory and Policy Initiatives 1-2 2.0 Issues in Compressor and Motor Selection 2-1 2.1 CompressorTechnology Options 2-3 3.0 Compressor Technology: Current State of the Art 3-1 3.1 Description ofCurrentCompressorTechnology.... 3-1 3.1.1 Reciprocating Compressors 3-1 3.1.2 Rotary Compressors 3-3 3.1.3 Motor 3-5 3.2 Performance 3-7 3.3 CompressorCosts 3-10 3.4 Losses 3-11 3.4.1 Motors 3-14 3.4.2 Losses Within the Pump 3-15 3.4.2.1 Mechanical Efficiency 3-15 3.4.2.2 Pressure Losses 3-16 3.4.2.3 Volumetric Efficiency 3-16 3.4.2.4 Suction Gas Superheat 3-17 4.0 Options for Improvement and Their Cost: Large Compressors 4-1 4.1 Reciprocating Compressors 4-1 4.1.1 Non-Lubricated LinearFree Piston Reciprocating Compressor 4-2 4.2 Rotary Compressors 4-2 5.0 Options for Improvement and Their Cost: Small Compressors 5-1 5.1 Reciprocating Compressors 5-1 5.2 Rotary Compressors 5-3 5.3 Estimated Cost ofImproved Efficiency Small Compressors 5-3 5.4 Approach to Scaling ofLosses 5-4 5.4.1 Reciprocating Compressors 5-5 5.4.2 Rotary Compressors 5-7 6.0 Potential Value of Improvements 7.0 Areas for R&D 8.0 Conclusions 9.0 References and Sources 6-1 7-1 8-1 9-1

iiiFigure 2-1: Figure 3-1: Figure 3-2: Figure 3-3: Figure 3-4: Figure 3-5: Figure 3-6: Figure 3-7: Figure 3-8: Figure 3-9: Figure 3-10: Figure 3-11: Figure 4-1: Figure 4-2: Figure 5-1: Table of Figures Lorenz-MeutznerRefrigerator/FreezerCycle 2-1 Typical Small Hermetic Reciprocating Motor-Compressor 3-2 Typical Small Hermetic Rotary Motor-Compressor 3-4 CompressorEfficiencies versus Capacity..................................... 3-10 Estimated OEM Price ofAverage Efficiency Refrigeration Compressors 3-11 Ideal PV Processes at Standard Conditions 3-12 ConceptualMeans ofIsothermalizing the Compression Process 3-13 R12 Pressure/Enthalpy Diagram Effect ofSuperheat 3-18 CompressorHeatDistribution 3-19 Low Side Compressor 3-20 Effect ofSuction Gas Temperature on CompressorWorkfor ConstantMechanical & CylinderEfficiency 3-20 Low Side versus High Side Crankcase in a Single Cylinder Compressor 3-22 Improvementin Efficiency of"Large" Compressors for Domestic Refrigerator/Freezers overthe PastDecade 4-1 Estimated Incremental CostofImproved Efficiency Large Compressors 4-4 Estimated Incremental CostofImprovedEfficiency Small Compressors 5-4

ivTable 2-1: Table 3-1: Table 3-2:
Table 3-3:
Table 3-4: Table 3-5: Table 4-1:
Table 4-2:
Table 5-1:
Table 5-2:
Table 5-3: Table 5-4: Table 6-1: Table 6-2: Table of Tables Alternative Refrigerants 2-3 Refrigerator/Freezer Manufacturers and Compressor Suppliers 3-8 Partial List ofCurrent High Efficiency Refrigerator/Freezer Compressors 3-9 EER Limiting Cases, Based on Alternate Compression Processes

........................................................................................................ 3-13 Typical CompressorInputPowerDistribution 3-14 MotorEfficiencies 3-15 "Large" Reciprocating CompressorInputPowerDistribution vs Efficiency Level 4-2 "Large" Rotary Compressor Input Power Distribution vs Efficiency Level 4-3 Estimated Small Reciprocating CompressorLimiting Efficiency and Input Power Distribution, Based on Loss Scaling 5-2 Small (200 Btu/hrNominal Capacity) Reciprocating Compressor InputPowerDistribution at Several Efficiency Levels 5-2 Estimated Small Rotary CompressorLimiting Efficiency and Input Power Distribution, Based on Loss Scaling 5-3 Scaling ofLosses in Reciprocating and Toary Compressor 5-5 Energy Savings 6-1 Present Value ofEnergy Savings Obtained with Increased Efficiency Compressors 6-3

vAcknowledgements

During the course ofthis work, anumber ofcompressor, motor, and variable speed drive manufacturers were contacted, whose assistance and cooperation we acknowledge and appreciate:

• Compressors: Tecumseh Products Co., Copeland, Americold, Danfoss, Embraco, Sanyo, and Matsushita

• Motors: General Electric, A. O. Smith, Emerson, and Baldor

• Variable speed drives: Toshiba, Emerson, Mitsubishi, Hitachi, Westinghouse, Magnetek, Lenze, Vee Are, Ranco, Inland, PMI, Minarik, EG&G, Fasco, Boston Gear, and Graham

vi1.0 Introduction

The design ofthe domestic refrigerator-freezer will be undergoing a significant set of changes overthe next several years, driven by interrelated developments in a number of national energy and environmental policy areas - the CFC phase out, growing concerns about global warming, DOE appliance energy efficiency standard setting, and others. These developments have created the needfor acomprehensive examination ofthe design options for domestic refrigerator/freezers. These interrelated options include compressors, motors, refrigerants and lubricants, the refrigeration cycle, cabinet insulation, and other aspects ofcabinetthermal design. This reportcovers the technology status ofthe small hermetic compressors used in the refrigeration system ofdomestic refrigerators, addressing the potential for and costof improvements in the efficiency ofsmall hermetic refrigeration compressors. The environmental and energy policy background is discussed briefly below.
1.1 Stratospheric Ozone Depletion and the CFC Phase Out Evidence accumulated over the past 15 years indicates thatfully halogenated chlorofluorocarbons (CFCs) have caused measurable deterioration ofthe atmosphere's stratospheric ozone layer, which plays a significantrole in attenuating solarultraviolet radiation. Increased levels ofultravioletradiation would have alarge number of undesirable effects, including increased levels ofskin cancers. Overthe pastfew years, this subject area has received renewed attention as the resultofobservations in the mid-1980s of "gaps" in the ozone layerin the vicinity ofthe poles. As aresult ofthis attention, the Montreal CFC protocols were concluded in Fall, 1987. This international agreement was signed and ratified by the majorfree world industrial nations requiring a freeze, then phased productioncurtailments, ofCFCs. By 1998, production ofCFC-11, CFC-12, CFC-1l3, CFC-1l4, and CFC-1l5, as well as certain "halons" were to be reduced to 50% of 1986 levels. An additional provision provided for periodic review of scientific evidence and adjustment ofallowable levels ofproduction accordingly. The reassessmentcompleted in 1990 resulted in a nearly total phase out ofCFCs by the year 2000. The Clean Air Act of 1990 has codified this accelerated CFC phase out schedule into U.S. environmental law. In April of 1991, NASA reported the results of satellite-based measurements ofstratospheric ozone levels indicating that ozone depletion of5% overthe mid latitudes has already occurred. The result ofthis development has been further acceleration ofthe timetable for CFC phase out. In November, 1992, the Montreal Protocol Copenhagen Amendments accelerated the complete phase out ofCFCs to January 1, 1996. This has a direct impact on R1F insulation and compressors which have been designed around the characteristics of CFC-ll and CFC-12, respectively.

1-11.2 Global Warming Concurrentwith the accumulation ofscientific evidence ofstratospheric ozone depletion, increasing concerns have been developing about global warming caused by increasing atmospheric concentrations ofcarbon dioxide and the trace greenhouse gases. The most significant ofthe trace greenhouse gases are the CFCs and methane. The fact that the CFCs are powerful greenhouse gases has reinforced the pressures to accelerate the CFC phase out time table. Measures to limit orreduce CO2 emissions have been proposed. Because CO2 is one ofthe basic combustion products ofall ofthe fossil fuels used to produce energy for heating, transportation, and electric power generation, measures to reduce CO2 emissions require the burning ofless fossil fuels. One regulatory measure to bring this about is increasing energy efficiency standards. 1.3 DOE Appliance Energy Efficiency Standard Setting In February, 1989, DOE issued a final rulemaking under the Energy Policy and Conservation Act, as amended, establishing minimum energy efficiency standards for most categories ofconsumer appliances (Federal Register, 1989a). Depending on the category, the standards take effectbetween 1990 and 1993. The standards for domestic refrigerators and freezers went into effect on January 1, 1990, generally requiring efficiency levels in line with the mostefficientproducts available in the late 1980s, whose efficiency was nearly double the levels prevailing only 10 to 15 years earlier. Global warming (and national energy security issues) have resulted in the recent adoption ofsignificantly reduced levels ofallowable electric energy consumption for all categories ofdomestic refrigerators and freezers, effective on January 1, 1993 (the 1993 standards reduce allowable energy consumption by approximately 30% from the levels under the current Federal regulations that tookeffecton January 1, 1990) (Federal Register, 1989b). 

### 1.4 Other RegUlatory and Policy Initiatives States are instituting reforms in planning and rate making that put demand reductions on an equal playing field with building additional supply capacity. Integrated resource planning, adopted by many states, requires utilities to evaluate every "resource" (demand reduction or supply) in terms oftotal societal cost. California, Oregon, Washington, most ofthe New England states, Wisconsin, and New York have adopted ratemaking processes in which utility rates or return on investmentis adjusted so that utilities do notlose profits for forgone Kwh sales, butcan profit from demand reductions. In California and several other states, the non-pollution aspect of demand reductions has led regulators to allow shared savings ofcustomerbill reductions to further increase utility profits. As a consequence ofthis change in utility regulation, the demand for efficient refrigerators is rising.

1-2The Golden Carrot/early retirementprogram is one concrete manifestation ofthis trend. Under the Golden Carrot, utilities are banding together to pool rebates to produce an incentive for production ofa RlF thatis 30% better than DOE's 1993 standard in the 18 to 22 cubic foot range. With the impetus described earlierfrom CFCs, global warming, and other state regulatory reforms, the Golden Carnot will provide a strong incentive for vastimprovements in energy efficiency.

1-32.0 Issues in Compressor and Motor Selection

To meet the dual challenge ofdesigning for new blowing agents and refrigerants, while meeting much higherefficiency standards, significantchanges in the design ofboth the cabinet and the refrigeration cycle are underconsideration. There are many design options, and itis important to evaluate the potential ofeach option to contribute significant energy savings, without using CFCs and without increasing the costofthe refrigerator beyond the value added by the change (with respect to energy costs or utility) or decreasing the utility ofthe refrigerator to the consumer. Design options that are being pursued in currentR&D programs include:

.. Low thermal loss cabinets ("super-insulated boxes") reduce energy consumption by reducing the amount ofcooling that is needed. To utilize heatexchangers effectively and minimize cycling losses, the compressorcapacity should be reduced in proportion to the thermal load. However, sufficientcompressorcapacity may still be required to provide a sufficiently fast pulldown for food preservation.

.. Dual refrigeration loops (separate refrigeration systems for the refrigerator and freezer compartment) take advantage ofthe increased COP at the higherevaporator temperature that can provide the required cooling ofthe fresh food compartment.

.. The Lorenz cycle, shown schematically in Figure 2-1 uses anon-azeotropic refrigerant mixture with two evaporators and an interchanger to operate the fresh food compartment and the freezer at separate evaporator temperatures for higher efficiency. (Lorenz, 1975) Figure 2·1: Lorenz-Meutzner RefrigeratorlFreezer Cycle

Temp Condenser 2

't-------O'-----iFresh Food Evaporatorf-----, HX1

HX2

5 Freezer Evaporator Entropy



# .. Variable speed compressor operation can save energy by allowing continuous operation atlow capacity, eliminating cycling losses and allowing more efficient utilization ofheatexchangers. Overspeed operation can provide additional capacity for pulldown. The lattercharacteristic might be particularly advantageous with high performance cabinets.

2-1One consequence ofthese design changes is a decrease in the compressorcapacity required to best match a given size refrigerator. Improved insulation and cabinetdesign will reduce cooling loads, reducing required compressor capacity. Two compressor/two evaporator systems meet the freezer and fresh food compartmentheatloads with separate refrigeration systems, obtaining, in theory, significantimprovements in the efficiency with which the load in the fresh food compartmentis met (higherevaporating temperatures, no defrostcycle). Even less compressorcapacity is needed, especially in the fresh food compartment where nominal compressor capacities ofonly 200 Btu/hr may be needed. While the Lorenz cycle does notinherently resultin adrastic compressorcapacity reduction, othercompressorproblems, such as high starting torque requirements, have been observed. Variable speed compressors will tend to be smaller displacement, with pulldown requirements met by overspeed operation. Present commercially available low capacity compressors «600 Btu/hrnominal capacity) have very poorefficiencies, low enough in some cases to completely negate the gains obtained from the design options described above. To realize the efficiency benefits ofreduced loads and dual evaporator systems will require improved efficiency, lower capacity compressors. Regardless ofthe R/F cabinet and refrigeration cycle design approach taken, increases in the efficiency level thatis available in refrigerantcompressors will resultin proportional increases in the efficiency ofthe refrigeratorusing the compressor. A major issue for compressor design is the change in refrigerant from CFC-12 to alow ozone depletion, low global warming potential refrigerant. While CFC-12 has been shown to be a significantpart ofthe cause ofboth stratospheric ozone depletion and global warming, it is an excellent working tluid for domestic refrigerator/freezers. Ithas a favorable pressure-temperature relationship, good thermodynamic efficiency, stability, total miscibility with low ccst mineral oil, and moderate temperature rise with compression and is non-tlammable, non-toxic, and low cost. Alternate refrigerants that do notcontain chlorine orhave currently acceptable ozone depletion potentials do not possess identical attributes ofCFC-12. Thus, compressor modifications or new designs will be required to adaptto the characteristics ofa selected alternative refrigerant. Table 2-1 lists some ofthe potential alternative refrigerants and their status, including potential substitutes for CFC-ll, CFC-114, and CFC-502, as well as for CFC-12. The majorworking tluid options include the near drop in replacements for CFC-12 (i.e., those refrigerants having vapor pressure-temperature curves close to that ofCFC-12), lowervapor pressure refrigerants such as HCFC-124, and non-azeotropic refrigeration mixtures (NARMs). Lower vaporpressure refrigerants mightbe utilized in low capacity systems (design options described above), ifshown to result in higher efficiency oflow capacity compressors, by virtue ofthe larger displacementthatwould be needed. The Lorenz cycle would utilize a NARM.

2-2Table 2-1: Alternative Refrigerants Displaced Probable Description, Status, Comment Refrigerant Availability HCFC-22 CFC-12 Current Commercially available, widely used refrigerant. CFC-502 Contains chlorine, will be phased out under Montreal Protocol Copenhagen Amendments HFC-134a CFC-12 Current Commercially available, rapidly expanding production Ternary CFC-12 (w/HCFC-124) Available in limited amounts 1993 HFC-152a CFC-12 Current Commercially produced and sold in fairly small quantities, used primarily as a component in CFC-500 (26%) and as a component in aerosol propellant blends HFC-123 CFC-11 1990-1991 Toxicitytests have shown sufficient toxicityto set AEL at 10 ppm; commercially available HFC-124 CFC-114 1993-95 Co-product of HCFC-123 production. Long term toxicity testing started HFC-125 CFC-502 Available in Near-term availability in blends to replace CFC-502 limited amounts HCFC-141 b CFC-11 Current Commercial production began in July 1988. Toxicity testing underway. Possible use as a foam blowing agent HCFC-142b CFC-12, 114 Current Used in R221R142b blends NHs CFC-11 Current Commercially available. Widely used in industrial CFC-12 refrigeration sector. Toxic with low flammability CFC-502 Source: Arthur D. Little, 1993 The majorissues that need to be considered in adapting the compressor design to an alternate refrigerantinclude the displacement required to obtain the intended capacity, lubricant selection, and material compatibility, especially the motor winding insulation. In summary, higher efficiency compressors are needed, especially in smallercapacities. Motor technology is an important consideration, because increasing the efficiency ofthe compressormotoris a straightforward way to improve compressor efficiency. For**variabl~** speed compressors, the variable speed motor and electronic drive represent the major technology component and the major costdriver.
2.1 Compressor Technology Survey In view ofthe importance ofcompressor performance to the potential efficiency improvements ofvarious design altematives, this review ofthe technology ofsmall hermetic refrigeration compressors was undertaken as part ofa larger study to evaluate

2-3the options for maximizing the efficiency ofdomestic refrigerators. The compressor characteristics documented in the report are being used in the overall study as part ofthe input database for modeling and evaluating design options. This reportcovers the technology status ofsmall hermetic compressors used in the refrigeration system ofdomestic refrigerators, addressing the potential for and cost of improvements in the efficiency ofsmall hermetic refrigeration compressors, with particularemphasis on smaller «800 Btu/hr nominal capacity) compressors. This reportis intended to serve three functions: 1) description ofthe state-of-the-art of currentcompressors and the potential for future improvements; 2) summary of performance and cost data as inputto evaluations ofrefrigerator/freezer system design options; and 3) presentpreliminary results indicating the level ofRIP energy consumption reductions thatcan be obtained through the use ofhigh efficiency small compressors.

2-43.0 Compressor Technology: Current State of the Art

3.1 Description of Current Compressor Technology Two types ofcompressors are used in current domestic R/F units, in the United States and worldwide:

.. Reciprocating

.. "Rotary" (Rolling Piston

- stationary vane) Both types are welded hermetic, i.e. the compressor pump and its motor are sealed inside a welded shelL The refrigerant gas connections are welded to the shell and the electrical connections to the motor are made through an insulated pass through bonded to the metal shell. This hermetic arrangement prevents the loss ofoil and refrigerant which could occurthrough rotating seals or mechanical fittings. 3.1.1 Reciprocating Compressors The reciprocating compressor has been and still is the mostcommon type used. Ithas reached a high state ofdevelopment, is mechanically efficient and reliable and is relatively less expensive to manufacture than its alternative, the rotary, because ofits lower overall sensitivity to manufacturing tolerances. A typical reciprocating compressor used in domestic R/Fs (Figure 3-1) is a single cylinder device with a piston driven by a crankshaftwhich is an integral extension ofthe driving motor shaft. The piston is connected to the crankby aconnecting rod and wrist pin. The piston reciprocates in a stationary cylinder secured to the motor stator. A cylinder head attached to the cylinder houses two reed valves, one ofwhich, the suction, opens into the cylinder and the other, the discharge opens outwardly from the cylinder. An oil pump is located in the non-driving end ofthe motor shaftwhich supplies oil to the rotating and reciprocating parts ofthe pump and motor. Gas is drawn from the R/F evaporatorinto the compressor shell in the space in the cannister surrounding the motor/pump combination. Itcirculates within the space aided by the fan effect ofthe motor rotor. This cools the motor and the pump. A significant portion ofthe heatpicked up by the gas is convected to the shell for dissipation to ambient air. The heated gas is then drawn through the pump suction valve into the cylinder where it is compressed and ejected through the discharge valve and piped to the discharge connection on the shelL Higherefficiency compressors may provide for a directed suction path from the inlet at the shell directly into a muffler assembly, reducing the superheating ofthe suction gas. Forced air cooling ofthe compressor shell may then be required to provide for motor cooling. The path ofthe suction gas ensures that oil which is discharged to the external system with the compressed gas is ultimately returned to the cannister and oil pump. It also allows liquid refrigerant which accumulates on shutdown in the evaporator to he slugged into the cannister on start up rather than into the cylinder. This could damage the suction valve on the compression stroke. The liquid is vaporized harmlessly by the

3-1Figure 3·1: Typical Small Hermetic Reciprocating Motor-Compressor

Suction Discharge / crankShaft Discharge Nuffler Cavity

.. 

### - M**\.l-IIIU~.11** HIT''+IIH'II"
-

## '=="

I ;: ....... 1-1111-1--,+-1-+-+-+-4·-

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# .-

..nl Piston & Connecting Rod Discharge Valve & Port

Suction Valve &Port

Electrical [ Conne"or

Mounting Springs Motor Rotor OJ.l Pump Inlet '- Oil Sump

Source: ADL sketch of Danfoss IL3A Compressor (315 Btu/hr, 3.5 EER)

3-2heat dissipated into the suction gas and returned to the cycle (a small energy penalty, contributing to cycling losses, is incurred due to the compression work associated with liquid being vaporized in the crank case). The pump/motor assembly is mounted in the shell with isolation springs to minimize vibration transfer to the shell. Mufflers contained within the shell as apart ofthe pump suction and discharge assemblies dampen gas pulsations. 3.1.2 RotaryCompressors This class ofcompressors for RJF use in the range ofinterest (200 to 800 Btus) is limited to the rolling piston

- stationary vane type rotary compressor. Rotary vane compressors were used in the pastbut are less efficientthan eitherthe reciprocating or rolling piston compressors and are no longeremployed. The rolling piston compressor (Figure 3-2) consists ofa cylindrical "piston" rolling on the wall ofa "cylinder" which is capped atboth ends. The piston is driven in the circular orbit defined by its rolling on the cylinder wall, by an eccentric which is an integral extension ofthe motor shaft. Areciprocating vane located in the cylinder between the discharge and suction ports is held against the piston by a spring to provide a seal between the suction and compression sides ofthe piston. Areed valve is used on the discharge port to prevent backflow from the discharge line into the compression cavity and short circuiting from the discharge to the suction during the interval when the piston is between the discharge and suction ports. The former ensures variable pressure compression. Rotary compressors used in domestic RJFs contain the motor and pump housed in awelded shellhermetically sealed against refrigerant leakage. Suction gas is piped directly to the suction port which minimizes superheating. The suction gas is not used to cool the motor/pump combination. Because ofthe close clearances, rotary compressors are very sensitive to ingestion ofboth particulates and liquid slugs. To protect againstboth ofthese, all rotary compressors include a suction accumulator and strainerclose coupled to the inlet. Discharge gas is ported directly into the shell which is at discharge pressure (high side) where entrained oil has an opportunity to partially separate before the gas exits the discharge connection. A portion ofthe heat dissipated in the motor is transferred to the discharge gas. The motor statoris pressed into the shell which promotes conduction of heat to the shell. Oilis metered to the bearings from the sump through an orifice by the pressure difference from the shell side and the crank/piston bearing area which is at a pressure intermediate to the suction and discharge pressure. The piston does not actually contactthe cylinderwalls orend caps, butis separated from them by a thin oil film. Close clearances are necessary to prevent gas blow by. Sealing ofthe clearances is enhanced by lubricating oil filling the clearance. Rotary compressors generally have a suction accumulator/filterclosely coupled to the inlet of the pump to absorb liquid refrigerant and oil slugs at startup and preventparticulates from entering the close clearances ofthe pump.

3-3Figure 3-2: Typical Small Hermetic Rotary Motor-Compressor

Rolling "Piston Crankshaft Eccentric
Oil Inlet To Shaft Port Motor Rotor
Motor Stator &Windings

# Illflll~l-suction Port

Sliding Valve and Spring

Suction Port

Source: ADL sketch of GE 2T/27RV/RS-36 Compressor (approximately 1000 Btu/hr)

3-4The rotary pump has less vibration than the reciprocating pump. Its piston has an eccentric orbit and the vane reciprocates; the eccentrically orbiting rolling piston is dynamically balanced by a counterweight and trim weight on the motor shaft. The reciprocating motion ofthe vane is notcounterbalanced, butthe productofthe weight and stroke length ofthe vane is atleast an order ofmagnitude less than thatofthe piston and connecting rod ofa reciprocating compressorhaving the same capacity. The noise level (vibration in the audible range) ofa rotary is higher than an equivalent reciprocating pump even though its total vibration is less, because the motor-pump assembly is rigidly attached to the compressor shell. The principal disadvantage ofthis pump is that itrequires close clearances between piston and cylinder, piston and crankshaft, crankshaft and supportbearings and piston and cylinder caps to preventexcessive gas blow by. These constraints require close machining tolerances (on the order offractions ofa thousandth ofan inch) over a much greater total surface area than the reciprocating pump with attendantcostpenalties and make the pump less tolerant ofwear. There is a general expectation that the rolling piston compressoris not going to be as efficient as a reciprocating pump in the lowercapacity ranges. This results from the inherently high surface to volume ratio ofthe cylinder and piston. This increases the leakage paths and decreases the mechanical efficiency at agreaterrate with reducing size than it does for a reciprocating compressor. This opinion was supported by Professor Soedel ofthe Herrick Laboratories at Purdue University (reference 15) and extended to scroll compressors particularly because ofthe disproportionate increase in leakage paths. The principal advantages ofthe rolling piston pump are thatitcan be a high side pump with attendant low suction gas superheat, high volumetric efficiency because ofits inherently low clearance volume and unrestricted suction port. These advantages in energy efficiency are offset by higher levels ofotherlosses, resulting in overall efficiencies being less than the bestcurrentreciprocating pumps in the highercapacities. Rotaries are more compact and lighterin weight than recips ofcomparable capacity; the inherentpotential cost savings have been a majorreason for their growth in market share during the 

### 1980's. 3.1.3 Motor Both types ofhermetic compressors are driven by single phase squirrel cage induction motors. The motors consist ofpairs ofstationary (stator) electromagnets placed 180 degrees apart, energized by "run" windings. The currentproduces an oscillating magnetic field in the plane centered on the motor axis. This field induces acurrentin rotor conductors which in turn generates a magnetic field in the rotoriron core. The rotor field opposes the statorfield producing atorque in the rotor.

3-5An induction motor depends upon the rotormagnetic field being generated by the induced rotorcurrent. This currentis induced only when there is a speed difference between the oscillating stator and rotating rotor fields. The speed ofthe statorfield is determined by the alternating currentfrequency and the number ofstatorcoil pairs; e.g., a 60 hertz currentwith atwo coil (pole) statorproduces a magnetic field oscillating at 3600 rpm; afour pole motor atthe same frequency produces an 1800 rpm field speed. The motor speed is the speed ofthe rotor, which must rotate at a speed less than that ofthe statorfield. A speed difference, or "slip", of approximately 3%, corresponds to the rated torque and poweroutput ofthe motor. For a two pole motor this speed is about 3500 rpm. Both the stator and rotormagnets are made oflaminated iron. The stator coils are generally ofwound copper wire. The rotorcoil is formed ofaluminum bars castinto passages in the iron core resembling asquirrel cage configuration, thus the name "squirrel cage" motor. Both magnetic fields induce eddy currents in the iron resulting in heating losses. These are minimized by laminating the iron. When the rotor is stationary, the oscillating stator magnetic field produces abalanced force in the rotor with no net torque, consequently some assistance is required to start a single phase motor. This is accomplished by apair ofelectromagnetic poles located 90 degrees from the "run" magnets called the "starting windings". On starting, these windings in conjunction with the "run" windings produce a rotating magnetic field which induces a net torque in the rotor. Once the rotorhas achieved a selfsustaining speed the start circuit is switched out ofthe circuit. A resistor or a capacitor is employed in series with the start winding to enhance the phase difference and thus the starting torque. The windings are switched out ofthe circuitby amechanical relay or apositive temperature coefficientresistor (PTCR). The latterupon heating increases in resistance to reduce the currentflow to negligible proportions. A motor with a starting winding only is called a splitphase motor. A motorwith a resistor in series with the startwinding is called a Resistor StartInduction Run (RSIR) motor. These types produce a non-uniform rotating field which cause an objectionable vibration in the rotor and mustbe switched out ofthe circuit after starting. Ifacapacitoris connected in series with the start winding a 90° phase difference between the run and start windings occurs resulting in a greater starting torque. This type ofmotor is called aCapacitor Start (CS) motor. A capacitor smallerthan the starting capacitorcan be inserted into the starting winding circuit and the starting winding then operated continuously. This results in a higher powerfactor which in tum reduces motorcurrent, decreasing winding resistance losses and improving motorefficiency by about 10%. This is called acapacitorrun (CR)

3-6motor. Ifthe motor also employs an additional capacitorin parallel with the "run" capacitorfor starting the combination is called a "Permanent SplitCapacitor" (PSC) motor. The"starting" capacitoris switched out ofthe circuit after starting. The winding resistance, eddy currents and windage friction represent the losses ofa motor. These produce heat which is transferred to the refrigerant gas and then to the shell for dissipation to the ambient air. Currently, larger (>700 Btu/hr), higherefficiency compressors for domestic R/Fs use PSC motors. Smallercompressors generally use lowercostRSIR motors. 3.2 Performance As discussed in Section 2, it is likely thatcompressors for future R/F designs will be smaller than those CUlTentemployed, approximately one halfto one third ofcurrent averages:

• Present capacity range - 500 to 1500 Btufhr

• Probable future capacity range - 200 to 800 Btu/hr A partial survey ofthe major U.S. refrigerator/freezer manufacturers and their compressor suppliers (Table 3-1) shows that a wide selection ofmodels in the 200 to 800 Btufhrrange is available to U.S. RtF manufacturers (Table 3-2). As shown by the data in Table 3-2 and graphically in Figure 3-3, an efficiency decrease parallels the decrease in size, as indicated by the EER* value. In Figure 3-3, several curves are shown to fit the individual EER data points. Somewhat arbitrarily, below 600 Btu/hr is labeled "small compressor", above is labeled "large compressor." CUlTentproduction state ofthe art represents acurve fit ofthe values of compressors that are in full scaleproduction and use by OEMs. Near term SOA is a curve fit ofthe bestEER values for compressors we could identify for which samples are being supplied to OEMs, with production likely to occur as needed for 1993 appliance production.

EER (Energy Efficiency Ratio) is defined as the compressor refrigeration effect in Btu per hour divided by the compressor motor input in Watts (combined units: BtulWatt-hr). This expression is a variation of the "Coefficient of Performance" (COP) but well adapted to the units in general use for expressing compressor data. The standard conditions for measuring the terms of this value are -10°F saturated suction temperature, 130°F saturated discharge temperature, gO°F liquid and suction vaportemperatures and 90·F ambient temperature.

3-7Table 3·1: RefrigeratorlFreezer Manufacturers and Compressor Suppliers

Whirlpool

hite/Frigidaire mana
Admiral/Maytag

Source: References 4, 24 Compressor Suppliers General Electric Danfoss Tecumseh Panasonic (Matsushita) Embraco Aspera Americold Panasonic Sanyo Tecumseh Panasonic Sanyo Embraco Tecumseh

3-8C IF A f . H" hEft"

. P "I L" f C able 3-2: artla 1st 0 urrent Igi IClency e r1gerator reezer ompressors Manufacturer Cap Dspl EER Motor Cooling Type Sanyo (C030) 139 0.109 2.32 RSIR Static Recip Sanyo (C-M40L12C) 180 0.13 2.6 RSIR Static Recip Tecumseh (AZ1320D) 200 0.136 3.13 RSIR Static Recip Embraco (EMI 20ER) 220 0.139 3.6 RSIR Static Recip Panasonic (S070LKAA) 278 0.184 3.1 RSIR Static Recip Embraco (EM 30SC) 315 0.184 3.89 RSCR Static Recip Panasonic (S090LKAA) 349 0.217 3.5 RSIR Static Recip Embraco (EM 40SC) 420 0.23 4.3 RSIR Static Recip Americold (ST104) 440 0.217 4.7 PTCRICR Fan Recip Embraco (EMI 40ER) 440 0.23 4.4 RSIR Static Recip Panasonic (D112LRAA) 476 0.264 4.4 RSCR Static Recip Panasonic (DA43L67) 516 0.264 4.75 RSCR Static Recip Embraco (EM 55SC) 565 0.305 4.41 RSCR Static Recip Embraco (EM I 55ER) 580 0.305 4.5 RSIR Static Recip Americold (ST1 05) 580 0.277 5 PTCRICR Fan Recip Americold (HG 106-1 ) 621 0.277 5.35 PTCRICR Fan Recip Panasonic (DA51 L88R) 635 0.311 5 RSCR Static Recip Panasonic (FN40R80R) 651 0.248 4.46 RSCR Static Rotary Americold (ST106) 660 0.312 5.06 PTCRICR Fan Recip Americold (HG1 07-1) 708 0.312 5.4 PTCRICR Fan Recip Tecumseh (AE1370W) 740 0.421 4.6 PTCRICR Static Recip Panasonic (RA48L83R) 754 0.294 4.86 RSCR Fan Rotary Panasonic (DA66L11 R) 806 0.402 5.15 RSCR Fan Recip Americold (ST107) 810 0.366 5.06 PTCRICR Fan Recip Americold (HG 108-1 ) 848 0.361 5.5 PTCRICR Fan Recip Panasonic (RA53L11 R) 850 0.323 4.8 RSCR Fan Rotary Americold (HG 109-1 ) 947 0.401 5.53 PTCRICR Fan Recip Panasonic (DA73L13R) 973 0.469 5.25 RSCR Fan Recip Panasonic (FN60R12R) 1000 0.372 4.61 RSCR Fan Rotary Americold (HG 110-1 ) 1029 0.473 5.51 PTCRICR Fan Recip Panasonic (FN70R16R) 1151 0.435 4.79 RSCR Fan Rotary Americold (HG111-1) 1193 0.5 5.55 PTCRICR Fan Recip T

CAP DSPL RECIP RSIR PTCR/CR RSCR PSC Capacity (Btu/hr) Displacement (cubic inches) Reciprocating Resistance Start Induction Run Positive Temperature Coefficient Resistor Start/Capacitor Run Resistance Start Capacitor Run Permanent Split Capacitor Source: Engineering data supplied by the manufacturerof each model

3-9Figure 3-3: Compressor Efficiencies versus Capacity 

## 6,.----------------------------------,

5.5 Current, State-of-the-Art Large**ReCiPs~** CFC·12 Refrigerant

______________________________________u • .__.. .A.**~--n---A--** _ A fA

5

4.5 

## ...

__ u •••

__ •• ...... ... __**n._~** n

__ .. __ n •• n -

_ III 

# ilL IiIl 

# • •

• IiII

_________ • ..... __ •

__ ._. • •• .. u

_ 

### e. Current, State-of-the-Art Large Rotaries
II: UJ 4 UJ

.----..---.------- ---------------------------...--._.------------.------------------ '-----------' 

# •

.. -

__ ..... _. _ ... - -

__ - -

__ - - - - • -

__ n n •• n

_ 3.5

3 

# • • Current, State-of·the-Art ·Small" Compressors Static Cooled • Fan Cooled, Recip

... Fan Cooled, Rotary

..

2.5 

# •

1200 1000 800 600 400 200 2'------......-------L-.-----.L------'-------L.-- -'--_--'

o Cooling Capacity, Btulhr
3.3 Compressor Costs When sold on an OEM basis, refrigeration compressorprices are negotiated between the compressor manufacturer and the refrigerator manufacturer. Compressormanufacturers consider negotiated OEM pricing to be commercially/competitively sensitive information, and are, therefore, reluctant to discuss the subjectin detail. Figure 3-4 plots the estimated OEM price level versus capacity corresponding to the "average efficiency" versus capacity plotted in Figure 3-1, based on general discussions ofOEM price levels with several manufacturers.

3-10Figure 3-4: Estimated Price of Average Efficiency Refrigeration Compressors

40 .--------------------------------...,

1,200 1,000 400 600 800 Nominal Capacity, Btu/hr 200 O'-------'-------'-------L----------J'-------'--- -'

o 30

10 20

3.4 Losses As discussed above, the maximum efficiency level thatis available in cUlTent, production compressors, in terms ofstandard rating conditions EER, is 5.0 BtulWatt-hr for "large" (nominal capacity above 700 Btufhr) and 3.5 for the small (200 Btufhr) compressors ofspecific interestto this study. The 5.0 EERoflarge compressors represents a significantimprovement overthe efficiency levels that were available in 1980; much smallerimprovements were made to the smallercompressors overthis time period. To place the following discussion oflosses, and the discussion ofpotential improvements (Sections 4 and 5) in perspective, itis instructiveto considerthe thermodynamic limits on compressorEER. At the standard rating conditions, two thermodynamically limiting compression processes can be defined that are theoretically consistent with the rating conditions. Figure 3-5 plots the two processes, for CFC-12.

3-11Figure 3-5: Ideal PV Processes at Standard Conditions

250 ,.------------------------,

200 ..........................................................•..................................................................**.~** 150 0. I \ II\ \**-----~._-,---** --_.--._-----..------.._------------..._--------------------------.-------_.-----.-_.._---------.._----.....--- I\\

\\\\

__._._. ..**_._),.\~** .._ .._._. .. .... . 0.-

..-.._-----------.-.._---------..._--_.

"""'"

" 50**:::::~**

3 2.5 1.5 2 Specific Volume It Ilbm Curve A Curve B 0.5 oL.-----'-----'------'-----'-----'----..........l

o

CurveA: Reversible adiabatic (isentropic) compression from the suction conditions (_10°F saturation pressure of 19.19 psia for R12, superheated to 90°F) to the discharge pressure (130°F saturation pressure 195.71 psia for R12). Compressorefficiencies are most commonly quoted relative to this thermodynamically "ideal" case. Implicit in the selection ofan adiabatic process as the standard is the presumption that heatrejection directly from the compression process is impractical. Forthis process with CFC-12, the ideal EER is 9.26, at 90°F suction temperature _10°F saturated suctiontemperature, 130°F saturated discharge temperature, and 90°Fliquid and ambient airtemperature. Curve B: Reversible adiabatic compression from the suction conditions to 130°F, followed by reversible, isothermal compression to the discharge pressure. As indicated by the curve, isothermalization reduces the compression work; the resulting ideal EER is 9.8. Conceptually, isothermalization ofthe compression process to this extentmight be approached by routing partially condensed refrigerantfrom the condenserto passages in the cylinderbody and head, then returning the refrigerantto the condenser, as shown schematically in Figure 3-6. The difference in performance between these two cases is relatively modest. While the compression process ofdomestic refrigerator compressors is, in fact, not at all adiabatic, with internal heattransfer and heatrejection from the compressor shell to ambient air having a significanteffect on the process, these heattransfer processes resultin suction gas heating, as well as isothermalization. Forthe purposes ofthe discussions inthis section and Sections 4 and 5, isentropic compression is taken as the "ideal" process. Figure 3-5 was developed for CFC-l2, the refrigerantthatis the basis for all current

3-12Condenser Figure 3-6: Conceptual Means of Isothermalizing the Compression Process
Cooling

# Pass~s )

# ~~~ilillllil~ ........... I > I (
S",'on

_ -

# I'---..:~~c~om~pr~e~ss~o~r**~~~~~{~~_** -_ -_

-_

-_

# ~jl--" ( Gas
- Compo Discharge 

# r

in Gas 

### I I Liquid to capillary tube/interchanger Motor

compressors. Table 3-3 summarizes the ideal EER for CFC-12, HFC-134a, HFC-152a, and cyclopropane; at standard conditions the difference in ideal EER between these t1uids is small. Table 3-3: EER Limiting Cases, Based on Alternate Compression Processes Ideal EER' Refrigerant Reversible Adiabatic Isothermar 130°F CFC-12 9.26 9.8 HFC-134a 9.32 9.8 HFC-152a 9.26 9.8 Cyclopropane 9.15 9.7 At standard rating conditions, 100% efficient motor and thermodynamically reversible compression process.

.. Curve B in Figure 4-1 Table 3-4 presents, for several compressorconfigurations, an approximate breakdown of the electric input power among the basic categories oflosses and useful (basis: reversible adiabatic compression) refrigerant vaporcompression. The "high efficiency reciprocating" compressor represents the highestefficiency compressors that are currently available in production, in nominal capacities above 750 Btu/hr. Compressors having this level ofefficiency have been in commercial production for the past 2 or 3 years, and are the result ofdevelopment efforts that were initiated in the late 1970s. The "small, low efficiency reciprocating" compressoris representative ofcurrent compressors whose nominal capacity is less than 400 Btu/hr. The "typical rotary" is representative ofboth Japanese and GE rotary compressors, available in nominal capacities greater than 600 Btu/hr. The following subsections discuss the individual loss mechanisms in greater detail. Itappears that by employing techniques that are now

3-13used in larger compressors, efficiencies ofsmallercompressors can be increased significantly. Sections 4.0 and 5.0 discuss the potential for improvementfor "large" and "small" compressors, respectively. Table 3-4: Typical Compressor Input Power Distribution High Efficiency Small, low Typical Rotary Reciprocating Efficiency (EER

## = 4.7) (EER

## = 5.0) Reciprocating (EER

## = 3.2) Input Electric Power 100% 100% 100% Motor Loss 17% 28% 22% Mechanical Losses 8% 10% 10% Suction Gas Heating 12% 16% 3% Discharge Port Losses 2% 2% 4% Low Side Pressure Losses 4% 4% 2% Compression and Expansion 2% 4% 3% Losses Piston Blow By/Internal Leakage 1% 1% 5% Power to Compressor
- Gas 54% 35% 51% Delivered·· 3.4.1 Motors Table 3-5 shows the efficiencies currently attained with compressor motors now in use for the range ofinterest. It also shows a significant deterioration ofefficiency that parallels size. In this case, the reasons are largely due to less attention to performance in smallermotors in the interest oflow costs rather than an inherent trend to reduced efficiency. GE, for example, says thatitis not currently developing higher efficiency motors for compressors below 600 Btu/hrbecause oflack ofdemand. Conversely, because European and Japanese RJFs are smallerthan US models, there is more emphasis on efficiency in smaller motors. In the US, aluminum wire is used in small motor windings, whereas the Europeans and Japanese use more efficientcopper.
The possible efficiencies in Table 3-5 can be attained by the following improvements:

• All motors operated as Permanent SplitCapacitormotors,
• Low eddy current and hysterisis loss steel through betterchemistry and annealing,

• Additional material (larger motors), i.e., more copper, bigger magnets, and

• Different types oflaminations in some motor designs.

Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet conditions to discharge pressure.

3-14Table 3-5: Motor Efficiencies Compressor Motor Today's Motor Possible Rating HP Type Efficiency Type Efficiency 200 Btu 1/16 RSIR 70-73% PSC 84% 400 Btu 1/8 RSIR 73-76% PSC 85% 600 Btu 1/6 RSIR&PSC 78-82% PSC 86% 800 Btu 1/4 PSC 80-84% PSC 86% Legend RSIR Resistance Start Induction Run PSC Permanent Split Capacitor These improvements could result in motor costs from 1.5 to 3 times present costs, an increase over current motor costs ofapproximately $10 to $25, according to General Electric sources. The largestincreases would be in the lower outputmotors used with the smaller capacity compressors. These costincreases, while substantial, would still leave the cost ofsmaller motors less than that ofthe currentlarger, more efficient sizes. The projected efficiencies show an essentially level curve for the range ofinterest. This effect alone would playa major role in flattening the compressorEERversus capacity curve. A separate report on the technology ofboth constant and variable speed motors (reference 3) has been prepared under this project. 3.4.2 Losses Within the Pump The major losses in the pumps (i.e., the compressor less the motor) are mechanical friction losses, including bearings and cylinder/piston friction, pressure losses through mufflers, valve ports and reeds, other suction pressure losses, and suction gas superheating. 3.4.2.1 MechanicalEfficiency. Bearings comprise the majorelements that determine the mechanical efficiency ofa pump. They consistofjournal sliding and bearings. Additional losses that might be categorized as mechanical losses include oil pumping and windage, which together amount to 1 to 2Watts. Journal bearings are sleeve types and consist ofmainjournal bearings on the crankshaft for both types ofpumps, connecting rod/journal and piston/journal bearings for reciprocating and rotary pumps respectively and wristpin/piston/connecting rod bearings for reciprocating pumps. As a general rule, to minimize manufacturing costs, refrigeration compressors have been designed as "members" ofa "family" of compressors covering a reasonably wide range ofcapacities, with as many common parts as practical. Common parts include motor rotor and statorlaminations, cylinder housing castings, crankshafts, and connecting rod. A design that is mechanically

3-15optimized for the loads involved in the larger capacity models ofthe family will have larger than optimum bearings, for example, in the smallercapacity models, and, therefore, somewhatlowermechanical efficiency. Sliding surface friction (with a lubricating oil film between parts) occurs between the piston and cylinder ofboth reciprocating and rotary compressors. The sliding bearing surface friction between the closely fitted cylinder and piston in arotary pump is quite high and accounts for the difference between the mechanical efficiencies ofthe two types ofpumps. 3.4.2.2 Pressure Losses. Low side pressure losses notonly decrease the efficiency of the compressor, butreduce the gas density in the cylinder at the end ofthe suction stroke, reducing volumetric efficiency and cooling capacity. Majorrestrictions in the inlet gas flow path ofreciprocating compressors include the suction muffler and the suction valve port/reed. The inletto arotary compressoris virtually unrestricted. The neteffect is to cause a slightly higher compression ratio than the external suction and discharge conditions dictate. Analogous pressure losses apply to the discharge side ofthe compressor, and result in some overcompression. The reciprocating pump has both a suction and discharge valve plate and smallerports while the rotary pump has only a discharge valve plate and large ports. These distinctions accountfor some ofthe relative differences in efficiencies. Increasing port sizes and valve plate sizes would benefitreciprocating pumps marginally. The value of increased discharge valve portarea in terms ofreduced discharge pressure loss mustbe balanced against the increased clearance volume ofthe discharge port. Rotary pumps, however, have little room for improvementin this respect. 3.4.2.3 Volumetric Efficiency. Volumetric efficiency is the ratio ofthe actual mass flow rate compressed by the compressor to the ideal mass flow rate based on the displacement and RPM and the suction gas density atthe inletto the compressor shell. The major losses to volumetric efficiency are

• Pressure losses between the inletto the shell and the compressorcylinder (see brief discussion above),

• Suction gas superheating (see below). The combined effect ofpressure losses and suction gas superheating is to reduce the density ofthe refrigerant vaporin the cylinder when the piston is at bottom dead center atthe end ofthe suction stroke, and

• Clearance gas reexpansion. The clearance volume is comprised ofafinite space between the piston and cylinder at the end ofthe stroke and the discharge valve port. The piston clearanceis necessary to allow for the build up ofmanufacturing tolerances ofthe piston, cylinderhousing, crankshaft, and connecting rod as well as differential thermal expansion and insures

3-16

# l

t iJ



### 9**~. '~**  againstthe piston mechanically striking the cylinder head and the suction valve in reciprocating pumps. The rotary pump does not have a clearance volume between the piston and cylinder and does not have a suction valve. The only clearance volume is a discharge port and the radius on the end ofthe vane, on the discharge port side ofthe vane. The clearance volume retains a portion ofthe compressed gas which is not displaced from the cylinder. This retained gas re-expands on the suction stroke and occupies space in the cylinder preventing the new charge ofgas being drawn in from fully occupying the cylinder. The effect ofthis phenomenon is to reduce the effective capacity ofthe pump, a corollary ofwhich is that the pump must have a larger displacement than ifits volumetric efficiency were 100%. Volumetrjc efficiencies ofreciprocating pumps tend to be in the 60% range, consistent with the not insignificant magnitude in a reciprocating compressor ofall three ofthe above listed basic volumetric efficiency loss mechanisms. In contrast, the volumetric efficiency ofrotary compressors is typically over 90%, because the design minimizes the basic losses, i.e., the direct suction minimizes both the heating ofthe suction gas and low side pressure losses and, as noted above, the clearance volume is minimal. In principle, the volumetric efficiency has an effect only on pump capacity in that the compression and re-expansion ofthe clearance gas is a reversible process. In reality, the clearance gas compression and reexpansion processes are notreversible, primarily because ofcyclic heat transfer between the gas and the cylinder walls and head and piston (the reexpanding clearance gas is somewhatcooler than during the compression process, and returns less work to the piston than was originally required to compress the clearance gas). Fmther, there is an indirecteffect on work due to the resulting increased displacement which has a greater absolute friction loss than a smaller one. In experimental work where clearance volumes were reduced from nominal values to the minimum value possible with a given set ofparts, the power input increased (because less work was returned to the piston from reexpansion ofthe clearance gas), but the refrigerant mass now rate increased more in percentage terms, giving an overall improvement in compressor EER. Forexample, experimental work by Westinghouse reported in 1981 showed that a 760 Btu/hrcompressor with a ratio ofsuction volume to clearance volume of83 and an EER of3.4 when equipped with a modified piston, had a 9% efficiency improvement. The suction volume to clearance volume ratio was increased to 187, the pump capacity increased to 924 Btu/hr and the EER improved to 3.7. 3.4.2.4 Suction Gas Superheat. To the extent that the refrigerant gas entering the pump cylinder is above the temperature ofthe gas at the inlet to the compressor shell, the work ofcompression will increase. This is shown, in somewhat oversimplified fashion, by the pressure enthalpy diagram (Figure 3-7) which shows that the compression work required to pump a gas through a given pressure difference increases with the suction gas temperature in the cylinder at the beginning ofcompression. (Note

3-17that both isentropic compression curves overstate the actual discharge temperatures that would occurin acompressorhaving the indicated cylinder gas temperature, butis generally illustrative ofthe effecton compression work ofincreased suction gas superheat.)
Figure 3-7: R12 Pressure/Enthalpy Diagram Effect of Superheat
-- 90"F versus 230"F (-10"F Saturated Suction, 130"F Saturated Discharge, Isentropic Compression)

000 ODO 400 300
2IlO

'00 '"' '"'
40 3D
20

10 0.0 6.0
4.0 10 2.0

1.0 0.0 0.6 0.4 0.3
0.2

0.1 ·20 I I I I I I I I I I I I I I I I I I I -I 6h

### = 29Btullb

20 40 60 10 110 510 100 110 120 130 140 150

Some superheatis desirable to prevent wetcompression and for certain other thermodynamic considerations; however, the suction gas is usually considerably superheated as a result ofambientheating and subcooling ofthe liquid refrigerant for increased refrigeration capacity. Figure 3-8 schematically shows the distribution ofthe sources ofsuction gas superheat for reciprocating and rotary compressors.
In reciprocating compressors, the suction gas gains heatfrom a number ofidentifiable sources. In atypical compressor, the suction gas enters the shell and mixes with gas that has been heated by the motor and otherhigher temperature internal surfaces, primarily the cylinderhead and cylinderbody and the discharge line. Next, the gas passes through the suction muffler, which is often a chamber orpair ofchambers bored into the (high operating temperature) cylinderhousing. Then the gas passes into the suction

3-18Figure 3-8: Compressor Heat Distribution

Suction--.

Gas Reciprocating Compressor**~~=----"IO>II=-** Discharge ,- Gas

a motor current Rotary Compressor 

# r--- Discharge a motor current 11 Gas

Oshell cooling

L---L Suction Gas

manifold in the cylinderhead, immediately adjacentto the high temperature discharge manifold. Finally the gas passes into the cylinder, whose walls have been 1:leated to a temperature intermediate to the inlet and discharge gas temperatures. As a result of these multiple heating steps, the suction gas superheatmay be as high as 150°Foverthe gas temperature at the shell connection. The suction line can be directly connected to the pump through the suction mufflerprovided that some type ofleakage path permits the pressure in the line to equalize with the shell gas. Figure 3-9 is an example ofhow this could be accomplished. This is being done in some larger R/F compressors. Matsushita reported a 6 to 10% improvementin compressorefficiency for a 25"F reduction in superheat at the pump portby using this approach. Figure 3-10 shows the theoretical increase ofcompressor work versus superheat over the gas inlet temperature to the shell, for the case ofa 90°F gas temperature to the compressor shell. Ifcylinder gas temperatures were reduced from theirtypical 200°F+ level to the inlet temperature at the compressor shell, efficiency would be improved by 20 to 25 percent. In practice, this is not achievable, because even after a direct suction connection through an insulated muffler is implemented, significant heat transferpaths remain in the cylinderhead and cylinder wall areas. The 5% to 10% efficiency improvement described in the preceding paragraph represents a practical level of improvement. In reciprocating compressors, as suction gas superheating is reduced, more heat must be rejected through the shell or through separate cooling ofthe oil. To reduce suction gas superheatfurther and lower pump temperatures, semi-hermetic designs could be utilized. In semi-hermetic and open drive compressors, pump heat is rejected through

3-19Figure 3-9: Low Side Compressor
Suction Muffler
Baffle Oil Outlet
Suction Chamber
Discharge Chamber Compressor Valve Head Joint Spring

Suction Tube

### r:-------------~

Suction Gas 

# L ,....--t-L--'----

Figure 3-10: Effect of Suction Gas Temperature on Compressor Workfor Constant Mechanical & Cylinder Efficiency

30 r--------------------------------.,.--,

oLJ.<:.- --l '--- .l- -'-- --'- --L.J 90 120 150 180 210 240 Cylinder Suction Gas Temp Degree F at Beginning of Compression 270

3-20the pump housing directly to the ambient air rather than through a gas interface to a shell and then to ambient air. The net result would be lowercylindertemperatures, further reducing cylinder gas temperatures. This would require an external suction muffler and would forego the sound isolation provided by the encapsulating shell, as well as hermetic sealing ofthe refrigerantcharge in the system. 

### In rotary compressors, the suction gas passes from the compressor shell directly to the cylinder inlet through a short (approximately I inch long) straightport in the cylinder body (see Figure 3-4). Even with heat transfer to the cylinder and rolling piston surfaces, suction gas superheatis very moderate in comparison with typical small welded hermetic reciprocating compressors, usually on the order of40° to 50°F. This low level ofsuperheatcoupled with the negligible suction portpressure loss accounts for the high volumetric efficiency that is typical ofrotaries. This level ofsuperheating is so low that there is notmuch potential for additional improvement. Unlike a rotary compressor whose crankshaft and motor operate at high side pressure, with the suction gas drawn directly into the pump, minimizing suction gas superheating, in a single cylinder reciprocating compressorthe crankcase orunderside ofthe piston operates at the suction pressure rather than the discharge pressure. The mechanical effect ofoperating a single cylindercompressor crankcase/piston underside athigh side pressure (to allow direct suction to the compressorcylinder) is to significantly increase the amount ofwork performed by the piston on each stroke (the theoretical network per complete crankshaft resolution is the same in eithercase), significantly increasing bearing loads and losses. As illustrated in Figure 3-11, with a low side crankcase, the theoretical pressure volume process follows the upper PV diagram. During the suction stroke, no pressure differential (neglecting losses) acts across the piston, and during the discharge stroke all ofthe PV work is performed. With ahigh side crankcase, as shown in the lower diagram, the refrigerant vapor in the cylinder on the top side ofthe piston still would follow the upper PV process. However, during the suction stroke, the differential between high and low side pressure would act across the piston. The resulting work done by the piston would be the area indicated in the lowerPV diagram, more than double the basic gas compression work. While this work would be "returned" to the piston during the discharge stroke, the work on each individual stroke would be handled mechanically; during the suction stroke, bearing loads, and losses would nearly be doubled.

3-21Figure 3-11: Low Side versus High Side Crankcase in a Single Cylinder Compressor

P2 P1 Discharge

Gas 

# -8- P2 P1 

## I Suction Gas Compressor Shell P1

P1 Suction Gas

(during suction stroke) P2 

# -8- I Compressor Shell P2 Discharge

Gas

3-224.0 Options for Improvement and Their Cost: Large Compressors

4.1 Reciprocating Compressors The "large" reciprocating compressors currently used in the majority ofU.S. manufactured domestic R/F have nominal capacities ranging between 600 Btu/hr and 1300 Btu/hr. In this capacity range, the maximum efficiency level ofproduction compressors has increased significantly overthe past decade, as indicated in Figure 4-1. In 1980, the maximum compressorEER in this capacity range was approximately 4.0 BtulWatt-hr. By 1992, available EERs had increased from 5.3 to 5.5. Currently, in early 1993, limited numbers ofcompressor samples having an EER of6.0 (using a high efficiency, electronically commutated permanentmagnet DC motor) are available. Figure 4-1: Improvement in Efficiency of "Large" Compressors for Domestic RefrigeratorlFreezers overthe Past Decade (EER of Isentropic Compressor with 100% efficient motor is 9.3 BtulWatt-hr) 

## 6r----------=====-------------, 1993 Samples

5.5

- ------- - --------..------------.••-.--..-----·"199XPrll-Protiactl'oIfSOA-----··--------------··-·---------

1990 Production SOA 5 -

..-------------...-.....-----........----..----------------------- ..--.

4.5

- •.----------- .....--...-------..----------------------------------------------.-.•.--.•--.--------••-..••----------.------.---. §. a: w w 198D Production SOA 4 -.--------------••-.----- ..... --•••-..--.-----.-.---.----------.-.....-...--- -----...--......----••--.--------

3.5

- ---...---------••-------....----••---.----------..-------.---------"

-------. --.-••--------------------------------.-------....

3 '----------'--,----

..... '--------','----------'---,-------' 600 800 1000 Cooling Capacity (Btulhr) 1200 1400

Table4-1 is a breakdown ofthe inputpowerfor large compressors atthe four efficiency levels shown in Figure 4-1. Significantreductions in motor losses and suction gas heating losses provided most ofthe pert'ormance improvementbetween the 4.0 EER level of 1980 and the 5.0 EER level ofthe bestcurrentproduction compressors. Further improvementto the 5.3 EER level has come from incremental reductions ofmechanical losses (primarily through the use ofareduced viscosity lubricant), motor losses, and suction gas heating. The OEM costpremium for this level ofcompressor efficiency is on the order of$5. EER's on the order of5.5 to 6.0 have been obtained through further incremental loss reduction, as indicated in the lastcolumn ofTable 4-1. The design

4-1measures include using the highest possible efficiency electric motor (electronically commuted permanentmagnet rotor DC motor) and reduced suction gas heating losses (probably requiring active cooling ofthe cylinderbody and head, as shown in Figure 3-6). This level is probably approaching the practical limits; the necessary design measures would add about $15 to the OEM price ofthe compressor. Table 4·1: "Large" Reciprocating Compressor Input Power Distribution vs. Efficiency Level 1980 Production Current Current Future Goal State of the Art Production Production (EER

## = 6.5) (EER

## = 4.0) (EER

## = 5.0) State of the Art Samples (EER

## = 5.3) Input Electric Power 100% 100% 100% 100% Motor Loss 23% 17% 16% 8% Mechanical Losses 9% 8% 7% 6% Suction Gas Heating 15% 12% 11% 8% Discharge Port Losses 2% 2% 2% 2% Low Side Pressure Losses 4% 4% 4% 3% Compression and Expansion 3% 2% 2% 2% Losses (clearance gas expansion) Piston Blow By/Internal Leakage 1% 1% 1% 1% I Power Delivered to Gas 43% 54% 57% 70% I Compression Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet conditions to discharge pressure. 4.1.1 Non-LubricatedLinearFree Piston Reciprocating Compressor In a project sponsored by the U.S. EPA, Sunpowerhas developed aprototype, non-lubricated linear free piston compressor. The piston is driven by an electronically-driven linear, permanentmagnetmotor. A measured EER at standard conditions of6.2 has been reported (References 40, 42). Efficiency improvements are attributed to high (94%) motorefficiency, having only minimal friction losses in the piston/cylinder gas bearing, offsetto some extentby additional piston blow by and some additional valve losses, both due to absence ofthe sealing effectofthe oil film, piston spring losses, and additional losses due to the unbalanced reciprocating motion ofthe piston. The claimed potential for further efficiency improvementhas notyetbeen demonstrated. Significantissues include the level ofreliability and operating life (40,000 to 80,000 hours withoutfailure in well over90% ofunits is routine in current compressors) that can be attained with an unlubricated free piston device and the costto manufacture, including necessary high energy permanent magnetmaterials and electronic drive in the motor. 4.2 Rotary Compressors In the "large" compressorcapacity range, currentproduction rotary compressors are approximately competitive with currentproduction reciprocating compressors with

4-2respect to both costand efficiency; within the range oftheir own manufacturing processes and cost/performance/reliability trade offs, different manufacturers have placed greater or lesseremphasis on each technology. Table4-2 summarizes the breakdown oflosses for currentproduction and future, improved efficiency "large" rotary compressors. The efficiency level ofthe middle column was reported in Reference (38), and was the result ofan intensive effortto reduce all ofthe major losses (except the motor losses) to the minimum possible level. A significantreduction in blow-by loss was achieved by reducing the rolling piston to cylinderclearances by afactor ofapproximately two. Incremental reduction in clearance gas re-expansion loss was obtained through reduction ofthe discharge port volume and an incremental reduction ofmechanical losses was achieved through optimization ofbearing diameters, lengths, and clearances. To date, no rotary compressoris in production at this EER level, perhaps being indicative ofthe difficulty in mass production ofholding the tolerances required to allow the reduced clearances needed to minimize blow by losses. Table 4-2: "Large" Rotary Compressor Input Power Distribution vs. Efficiency Level Typical Rotary Reference (38) With Best Motor

(EER = 4.7) (EER=5.1) (EER = 5.6) Input Electric Power 100% 100% 100% Motor Loss 22% 22% 16% Mechanical Losses 10% 8% 9% Suction Gas Heating 3% 3% 3% Discharge Port Losses 4% 31/2% 31/2% Low Side Pressure Losses 2% 2% 2% Compression and Expansion 3% 21/2% 21/2% Losses Piston Blow By/Internal Leakage 5% 4% 4% Power to Compressor
- Gas 51% 55% 60% Delivered*

Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet conditions to discharge pressure. The lastcolumn shows the efficiency level resulting from substituting a high efficiency motor for the motor used in the Reference (38) compressor. Due to the higher operating temperature and smaller diameter ofthe motor in arotary compressor, the best motor efficiency is approximately 2 percentage points less than that ofthe motorfor a comparable capacity reciprocating compressor. The resulting EER of5.6 is essentially comparable to the estimated limiting efficiency range for a "large" reciprocating compressor.

4-34.3 Cost of Increased Efficiency As discussed in Section 3.3, manufacturers are generally reluctant to discuss their OEM pricing policies in any detail, citing both domestic and foreign competition and confidential relationships with theircustomers. Figure 4-2 plots the estimated OEM costpremium ofincreased efficiency for large compressors. The plotis based on numerous informal discussions with compressor manufacturers and other industry participants. A significantportion ofthe costincrease is attributable to incremental increases in motorcosts, as the maximum feasible efficiency ofthe motor is approached. Figure 4-2: Estimated Incremental Cost of Improved Efficiency large Compressors

18

16

14

l!! 12 .!11 (5 0 of 0 10 U UJ 0 "iii 8 C Q) E
u 6 E Base EER =4.7
4

2

04.5 5 5.5 6

No attempthas been made to estimate the magnitude ofthe capital investments in R&D and new tooling that would be required to develop new, higherefficiency compressors.

4-45.0 Options for Improvement and Their Cost: Small Compressors

As indicated in the preceding sections, the efficiencies ofthe smallercompressors that are the focus ofthis study and will be needed to effectively utilize improved cabinet thermal design and dual evaporatorcycles currently lag the efficiency that is available in the best compressors utilized in largerrefrigerators and freezers. Where the efficiency level oflargercompressors has been improved substantially since 1980, only limited improvements in efficiency have been implemented in the smallestcompressors. In order to achieve EERs in the 200 to 600 Btu capacity range comparable to those currently available in the highercapacity machines efficiency improvements in both the motor and pump are required. In this section, the potential for improvements is discussed in terms of specific options for improving small reciprocating compressors and rotary compressors. The emphasis is on the 200 to 400 Btulhrcapacity range needed for dual loop systems. Owing to the lack ofcurrent production, high efficiency, small compressors, a set of loss scaling relationships have been developed to assess the extent to which the efficiency improvements obtained inlarger compressors mightbe obtained in small compressors through similar design modifications.
5.1 Reciprocating Compressors Table 5-1 is an estimate ofthe maximum efficiency level that is attainable in a small, 200 Btulhrnominal capacity reciprocating compressor, based on applying the loss scaling factors discussed in Section 5.4, to the maximum feasible efficiency large compressor (nominal capacity 800 Btulhr, EER 6.0) discussed in Section 4.1. The resulting EER of5.3 represents both a substantial improvement overcurrentproduction small compressors and a significantR&D challenge. Consistent with the upperlimits on small compressorperformance estimated in the preceding paragraph, Table 5-2 is a breakdown ofthe inputpowerin a nominal 200 Btulhrcapacity compressor atcurrentlow efficiency levels, at an intermediate efficiency level reached primarily through improved motorperformance, and at a maximum feasible efficiency level, where measures have been taken to reduce all losses. The intermediate efficiency level is reached through motor efficiency improvement. The maximum feasible efficiency level is achieved by using the highest efficiency motor and implementing measures to reduce mechanical losses, clearance volume and suction gas heating.

5-1Table 5-1: Estimated Small Reciprocating Compressor Limiting Efficiency and Input Power Distribution, Based on Loss Scaling large loss Scaling loss Scaling limiting Reciprocating Factor** Factor @ 1/4 Efficiency @

at limiting Capacity*** 200 Btu/hr Efficiency (EER = 6.5) Input Electric Power 100% N/A N/A 100% Motor Loss 8% Table 3
- 5 +2% 10% Mechanical Losses 6% Capacity -1/6 1.26 7.5% Suction Gas Heating 8% Capacity -1/6 1.26 10% Discharge Port Losses 2%


### -- 

## -- 2% Low Side Pressure Losses 3%


### -- 

## -- 3% Compression and Expansion 2% Capacity -1/3 1.6 3.2% Losses Piston Blow By/Internal Leakage 1%


### -- -- 1% Power Delivered to Gas 70% N/A N/A 63% Compression* EER 5.7

*** Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet conditions to discharge pressure. Scaling factors are discussed in 5.4 Comparison of 200 Btu/hr to 800 Btu/hr nominal capacities Table 5-2: Small (200 Btu/hr Nominal Capacity) Reciprocating Compressor Input Power Distribution at Several Efficiency levels low Efficiency, Current Intermediate Efficiency Maximum Efficiency Production SOA (EER = 4.2) EER =5.7 (EER = 3.2) Input Electric Power 100% 100% 100% Motor Loss 28% 18% 10% Mechanical Losses 10% 10% 8% Suction Gas Heating 16% 16% 10% Discharge Port Losses 2% 2% 2% Low Side Pressure 4% 4% 3% Losses Compression and 4% 4% 3% Expansion Losses Piston Blow By/Internal 1% 1% 1% Leakage Power Delivered to Gas 35% 45% 63% Compression*

Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet conditions to discharge pressure.

5-25.2 Rotary Compressors Table 5-3 is an estimate ofthe maximum efficiency levelthat is attainable in a small, 200 Btulhrnominal capacity rotary compressor, based on applying the loss scaling factors discussed in Section 5.4, to the maximum feasible efficiency large compressor (nominal capacity 800 Btulhr, EER 5.6) discussed in Section 4.2. The resulting EER of 4.4, which is considerably less than the estimated limiting efficiency for small reciprocating, is indicative ofthe extentto which scaling factors limit the efficiency potential ofrotary compressors for the smallestcapacity applications. Table 5-3: Estimated Small Rotary Compressor Limiting Efficiency and Input Power Distribution, Based on Loss Scaling "Large" Rotary Scaling Loss Scaling Small Rotary

at Limiting Factor·· Factor @ 1/4 Limiting Efficiency Capacity··· Efficiency @ (800 Btulhr, 200 Btulhr EER

## = 5.6) Input Electric Power 100% N/A N/A 100% Motor Loss 16% Table 3
- 5 +2% 18% Mechanical Losses 9% Capacity -1/6 1.26 12% Suction Gas Heating 3% Capacity ·1/3 1.6 5% Discharge Port Losses 31/2%


### -- 

## -- 31/2% Low Side Pressure Losses 2%


### -- 

## -- 2% Compression and Expansion 21/2%



### -- -- 21/2% Losses Piston Blow By/Internal Leakage 4% Capacity -2/3 2.5 10% Power Delivered to Gas 60% N/A N/A 47% Compression" EER 4.4 Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet conditions to discharge pressure. Scaling factors are discussed in 5.4 Comparison of 200 Btu/hr to 800 Btu/hr nominal capacities 5.3 Estimated Cost of Improved Efficiency Small Compressors As discussed in Section 3.3, manufacturers are generally reluctant to discuss their OEM pricing policies in any detail, citing both domestic and foreign competition and confidential relationships with theircustomers. Figure 5-1 plots the estimated OEM costpremium ofincreased efficiency for small reciprocating compressors. The plot is based on numerous informal discussions with compressor manufacturers and other industry participants. A significantportion ofthe costincrease is attributable to incremental increases in motor costs, as the maximum feasible efficiency ofthe motor is approached. No attempt has been made to estimate the magnitude ofthe capital investments in R&D and new tooling that would be required to develop new, higher efficiency compressors.

5-3Figure 5-1: Estimated Incremental Cost of Improved Efficiency Small Compressors

5.5 5 4.5 4 3.5 OL----'''''-----J'-- --' --' --L

_

3 20

I!! .!!! "0 Q 15 i 0 () ::ll w 0 iii C 10 

# .. E e u .E

5 25

5.4 Approach to Scaling of Losses To estimate the potential efficiency level thatcould be obtained in small compressors through the application ofthe same design features that have been used to improve large compressorefficiencies, a set ofsimple scaling relationships relating the magnitude of each loss to the capacity ofthe compressorwas developed, for both reciprocating and rotary compressors. Table 5-4 summarizes both sets ofloss scaling relationships, giving the variation with capacity ofthe loss as apercentage ofthe total powerinput. The assumptions and rationale for the treatment ofeach loss are discussed briefly below. These scaling relationships generally apply to the 200 Btu/hrto 800 Btu/hrnominal capacity range, at aconstant 2-pole motor speed ofapproximately 3500 RPM. A basic assumption underlying these scaling relationships is that the small capacity design is geometrically similarto the large capacity design, i.e. all dimensions are scaled by the same factor, with the exception ofclose clearances. To a first order, the capacity and inputpower are proportional to the displacement, which is proportional to the cube of the linear dimensions, e.g. diameter, length. This simplified treatmentignores many complexities and interrelations ofimportant design variables, such as the temperature dependence ofoil viscosity; as such it can be regarded in the aggregate as being generally indicative ofthe magnitude ofthe effect ofcompressorcapacity on the losses and the resulting efficiency potential.

5-4Table 5-4: Scaling of losses in Reciprocating and Rotary Compressor Scaling Rule for loss vs. Capacity Reciprocating Rotary Motor Loss Small
- Table 3-5 Small Table 3-5 Mechanical Losses Capacity -1/6 Capacity -1/6 Suction Gas Heating Capacity-1/6 Capacity -1/3 Discharge Port Losses No Scale Effect No Scale Effect Low Side Pressure Losses No Scale Effect No Scale Effect Clearance Gas Reexpansion Capacity -1/3 No Scale Effect Piston Blow By/Internal Leakage No Scale Effect Capacity -2/3 Motor losses. The treatment in Section 3.4.1 is the basis for relating losses to capacity, with Table 3-5 summarizing the relationship between capacity and motor efficiency. The maximum motor efficiency for arotary compressoris about 2percentage points less than the values indicated in Table 3-5, because ofthe higher operating temperature and larger aspectratio ofthe motor. High and Low Side Pressure Losses. Forboth recips and rotaries, the effect of geometrically scaling, atconstantrotational speed, vaporpassages and valve port diameters is to reduce gas velocities and pressure losses, complicatedby considerations such as the effect ofinlet and discharge valvereed stiffness on pressure drops. The overall effect ofscale is probably small, and is neglected in this treatment. This is the one area where the scaling factors are favorable to small capacities. 5.4.1 Reciprocating Compressors Mechanical Losses. Mechanical losses include shaft bearing friction losses and piston-cylinder wall sliding friction (hydrodynamic bearing oil film shear stress in both cases). Other, smallermechanical losses include oil pumping and windage.

• Piston friction powerdissipation is the product ofthe average oil film shearforce acting on the piston (oil viscosity x piston sliding surface areax average velocity + clearance) and the average piston velocity. The basic scaling relationships are: Average velocity oc diameter Piston sliding surface area oc diameter squared Oil viscosity is constant The diametral piston to wall clearance is constant, on the order of0.0001 to 0.0002 inches, dictated by the minimum practical level for selective assembly Overall, the power dissipation is proportional to d 4

, the ratio ofpower dissipated

. . d4 /d1 d . 1/3 to mput powerIS

. = = capacIty

5-5.. Shaftbearing friction: for a givenjournal bearing lubricantviscosity, LID ratio, clearance to diameterratio, and operating eccentricity, at constantrotational speed the power dissipated in the bearing is proportional to the design load, which varies with the cube ofthe shaftdiameter. The load, which originates primarily from gas pressure acting onthe piston is proportional to the piston diameter squared. The ratio ofthe powerloss (proportional to the bearing loads) to the total power inputis d2

/d3 = dol= capacity"1I3.

.. The net effect ofthe relatively small piston friction plus the bearing friction is estimated to be: mechanical powerlosses/total inputpower = capacity"1I6. Suction Gas Heating. As discussed below, underrotaries, suction gas heating varies with surface area to volume ratio, given aconstantinternal temperature distribution. In small recips, suction gas heating is a more significantloss, and affects the temperature distribution. With forced aircooling ofthe compressor shell, average internal temperatures will be reduced to a greaterextent at smallercapacities. Overall, alesser dependence ofsuction gas heating on scaleis assumed: capacityo 1l6. Clearance Gas Reexpansion. The magnitude ofthe ratio ofthis loss to total input power depends on the ratio ofthe clearance volume to the displacement. In larger, high efficiency compressors, this ratio has been reduced to the minimum practical level, and involves the handling ofpiston, cylinder, and connecting rod length tolerances, with a resulting minimum piston crown to cylinderhead clearance attop dead center. The minimum tolerance basedclearance will not scale down with reduced dimensions to any significant extent. The piston stroke does scale, so the ratio ofclearance volume to displacement will be proportional to dolor capacity to

0 113• Piston Blow-by. In arecip, the clearance between the piston and cylinderis small (diametral clearance between 0.0001 and 0.0002 inch) and filled with an oil film. The loss is generally less than 1% and scaling was neglected in this treatment.

5-65.4.2 RotaryCompressors Mechanical losses. Crankshaftbearing loads and losses and piston sliding losses scale in the same fashion as with recips. (overall loss to power outputratio scales with capacitil/ 6). Suction Gas Heating. This is a fairly small loss in a rotary, and consequently moderate changes in the loss will nothave a significanteffect on temperature levels throughout the compressor. The magnitude ofthe heat transferfrom the high side shell and the cylinder walls to the suction gas will be proportional to areas, or d2

. The ratio of suction gas heating to capacity is proportional to d 2 /d3

## =d-l

## =capacitil/3• Clearance Gas Reexpansion. The clearance volume in a rotary will scale with the overall dimensions; therefore no significant scale effectis expected. Piston Blow-By. This is a major loss in "large" rotaries, kept under control by rigorous control ofall internal clearances to the minimum practical level. It is doubtful whether significantly smallerclearances can be maintained in smallercapacity rotaries. Based on the assumption that the clearances are constant, the blow by loss is proportional to the perimeter which is proportional to d. The ratio ofthe loss to the capacity is d/d3 

## = d-2

## =capacity"2/3.

5-76.0 Potential Value of Improvements



### Table 6-1 summarizes the annual energy savings obtainable through improved compressorefficiencies. These savings have been evaluated for two basic R/F (top mounted freezer) configurations: single evaporator/single compressorand dual loop. The baseline, single evaporatorconfiguration is representative ofaconfiguration whose efficiency approximates the 1993 Federal Energy Efficiency Standards and includes the following basic design features: Table 6-1: Energy Savings

Large Compressor Single Evaporator

Small Compressors Dual Loop Compressor Efficiency EER Comment Annual En kWh* 4.85** Readily available 685 5.0 Best production 669 16 compressor 5.5 1993 production 618 67 6.0 Approaching practical 570 115 limits 3.6/4.7 Currently available 602 83 4.6/5.1 With efficient motor 537 148 5.3/5.5 Approaching practical 498 187 limits At DOE test conditions, taken to be representative of actual use ** Annual energy savings relative to large compressor/single evaporator baseline



### " Cabinetvolumes (cubic feet): freezer 4.6 fresh food 13.4 total 18.0 adjusted volume 20.9

" Federal energy efficiency standard: 1990 962 kWh/yr 1993 690 kWh/yr

" 4.85 EER, 893 Btulhrcompressor (readily available, moderate cost, reasonable efficiency, provides typical pulldown capacity).

" Cabinetinsulation is polyurethane foam with R=8°F/in perBtulhr-ft2 

### average freezer wall thickness: 23/8 inch average fresh food compartment wall thickness: 1 3/4inch insulation thickness in doors: 1 5/8 inch

" Standard evaporator (2 rows deep x 8rows high)

6-1• Efficient (PSC type) fans:

- evaporator 50 CFM, 6.8 Watts

- condenser90 CFM, 6.8 Watts

• Electric resistance anti-sweat heaters: 5.5 Watts for mullion (the energy analysis assumes that the heater poweris on one halfofthe time to approximate the DOE test procedure)

• Vapor-line anti-sweat heat for cabinet doorflange (cycle average of4 Watts)

• Defrostenergy is assumed to be constant across all cases studied. Variations in cycle time are assumed to be compensated by adjustments to the compressortimer interval. This baseline configuration is similarto the baseline configuration that was used in the MotorTechnology Study (Reference 3) to evaluate efficiency improvements obtainable from variable speed comprerssor and fan operation. The dual loop configuration uses the same cabinet as the single evaporatorconfiguration. The differences are:

• Evaporators

- each evaporatoris the same size and capacity ofthe single evaporator.

• Fans

- each evaporator and condenserfan is the same size as in the single evaporator cycle

• Compressors, currently available small compressors: freezer: 4.7 EER, nominal capacity is 440 Btu/hr

- fresh food: 3.6 EER, nominal capacity is 208 Btu/hr Foreach ofthese basic configurations, the annual energy consumption at the DOE test conditions was calculated for the baseline compressor efficiency, and for progressively increased compressorefficiency, as indicated in Table 6-1, consistent with the discussions in Sections 4 and 5. Forboth single evaporator and dual loop systems, the annual energy savings are relative to the baseline single evaporator configuration. The value ofthese energy savings is tabulated in Table 6-2 assuming an electric energy cost of8¢ perkWh, for a range ofreal (after inflation) discount rates between 2 and 10 percent, over the projected 15 yearlife ofthe refrigerator. The discountrates cover a range between real, after tax returns to savings accounts at the low end ofthe range, to after inflation creditcard interest rates at the high end ofthe range. Forreal discount rates in the range thatconsumers would rationally choose for safe investments (at the low end ofthe range), the present value ofthe saved energy is on the order of$50 to $100 greater than the incremental cost, at retail, ofthe increased compressorefficiency.

6-2Table 6-2: Present Value of Energy Savings Obtained with Increased Efficiency Compressors Present Value for Discount Cost of Rate Improvements Compressor Savings 2% 5% 10% (in U.S. Dollars EER kWh/yr** at Retail) Large Compressor 4.85*


### -- -- -- -- 0 Single Evaporator 5.0 16 17 14 10 5 5.5 67 70 58 43 10 6.0 115 120 99 73 20 Small Compressors 3.6/4.7 83 87 71 3 20 Dual Loop 4.6/5.1 148 155 127 94 40 5.3/5.5 187 196 161 169 80 Baseline
** Relative to baseline

6-37.0 Areas for R&D

As discussed in the preceding sections, considerable potential exists for improving the efficiency ofthe small hermetic compressors that would be used in dual loop designs. The areas for improvementinclude increasing motorefficiency and reducing losses in the compressor: suction gas superheat, clearance gas re-expansion, valve losses, and mechanical losses. The techniques for designing higher efficiency induction motors are well known in the hermetic motor industry. The majorissues involved from the industry viewpoint are productengineering and marketing issues ofcost vs. market size, and integration with the compressor. Beyond these commercialization issues, provision ofahigh efficiency motor for a small compressoris relatively straightforward. Because the motoris an integral part ofthe compressorpackage, high efficiency motor development is best carried out in the context ofhigh efficiency compressor development; integration ofa high efficiency motor with an existing small compressorrepresents alow risk, readily implementable means ofobtaining a significantincrease in small compressorefficiency. Early development ofsmall, increased efficiency compressors would provide important supportto currentprograms to demonstrate the performance potential of

### RIP's with dual loop systems and/or superinsulated cabinets. The focus ofa project should be the small, 200 Btu/hrnominal capacity (at standard rating conditions) compressorintended for the fresh food compartment ofa dual loop refrigerator/freezer oftypical (18 to 22 cubic feet) refrigerated volume. The results ofthe performance and payback analysis discussed in Section 6 clearly indicate that significant energy savings can be obtained with adual loop refrigeration system in arefrigerator/freezer, ifthe compressorefficiencies are the levels thatcan be attained with the improvements discussed in Sections 

### 4 and 

### 5, with the energy savings at current electricity prices providing a 3 to 5 year payback ofthe projected $40 to $80 increase (at retail) in compressorcost. To supportongoing efforts to develop and demonstrate this performance, early development ofa small, maximum efficiency compressorisjustified. As discussed in Section 5, the target performance for this compressor, at standard rating conditions (-10/130/90/90/90), would be:

• 5.0 EER

• 200 Btu/hr The basic approach would be to utilize the highest efficiency motor available and modify current small compressor designs to significantly reduce suction gas superheating. Otherloss areas should also be systematically examined for improvement potential. With the continuing acceleration ofthe CFC phase out, anon-CFC refrigerant and compatible lubricant should be utilized.

7-18.0 Conclusions

This study has dealtprimarily with the extentto which the efficiency ofsmall welded hermetic refrigeration compressors used in domestic refrigerators and freezers (having nominal capacities well below 800 Btu/hr) can be increased to levels approaching the state ofthe art (in production) for largercapacity welded hermetic domestic refrigeration compressors. The efficiency levels now attained in the compressors used in the majority ofdomestic refrigeration freezer applications (EERs between 4.8 and 5.3 BtulWatt-hr, for nominal capacities between 700 and 1100 Btu/hr) are the result ofdevelopmentefforts that were initiated in the late 1970s. The hermetic refrigeration compressors in this capacity range that were available then had EER values similarto today's small compressors. Developmentefforts have been focussed on achieving significantefficiency gains, but within very tight cost constraints and high reliability requirements. Mass scale production and specification in RJFs ofthe resulting high efficiency (but somewhat higher cost) compressors has been driven by a combination ofconsumer demand for higherefficiency refrigerators and the Federal energy efficiency standards. In response to continuing tightening ofthe energy efficiency standards, compressor manufacturers have continued to investin the development ofhigher efficiency, large compressors. Preproduction samples having an EER of5.3 to 5.5 became available to RlF manufacturers in 1991. Production ofthese compressors began in 1992. In anticipation oftighter standards in the future, there are continuing R&D programs aimed at developing still higherefficiency compressors. Samples ofECM motor driven compressors having an EER of6.0 have become available in early 1993. The practical, physical limit for the EER ofcompressors in this size range is approximately 6.3 to 6.5. To date, the small capacity compressors used in small, countertop refrigerators have not been involved in the efficiency upgrade programs because there has been no market pressure to improve the efficiency ofthese small refrigerators thatis comparable to marketcostpressures (retail prices for small countertop refrigerators typically range between $99 and $199; in relationship to these retail price levels, $10 to $20 to upgrade the efficiency ofthe compressoris a significantincrementin cost). Efficiency in the 200 to 600 Btu range ofcompressors can be improved with the application ofcurrenttechnology. Reciprocating compressors can be scaled down to low capacity levels with much less efficiency penalty than rotaries. Efficiency levels on the order of5.0 EER can be attained at the small end ofthis range through application of currenttechnology. The major design changes needed to reach this performance level are:

• Substitution ofhigherefficiency PSC or RSCR motors for the RSIR motors used in current small compressors,

8-1.. Implementing design features to reduce suction gas superheating in the compressor, including direct suction through an insulated muffler and forced aircooling ofthe compressor shell,

.. Reducing clearance volume at piston top dead center, to improve volumetric efficiency, and reduce clearance gas reexpansion losses, and

.. Reducing mechanical losses by optimizing the mechanical design for the bearing loads involved in the smallercapacities (instead ofthe larger compressors in the "family" ofcompressors ofwhich the small compressoris amember). The OEM costpremium ofan improved efficiency, 200 Btu/hr nominal capacity compressor would be on the order of$10 to $20 over the typical currentOEM cost of approximately $20 to $25. Regardless ofeconomic payback time, these improvements in small compressor efficiency may be anecessity for meeting the 1993 Federal energy efficiency standard for domestic refrigerators and freezers, or more stringent future standards. Two generic types ofcompressors are now in common use in welded hermetic compressors use in domestic refrigerators, reciprocating and rotary. Forthe smallest required capacities, e.g., for the fresh food compartment ofa dual evaporator system, reciprocating compressors are the more promising option for attaining high efficiencies.

8-29.0 References and Sources



### 1. Federal Register, April 21, 1989a, pp. 6077-8. 2. Federal Register, November 17, 1989b, pp. 47916-47938. 3. Dieckmann, J. and McMahon, E., 

### State ofthe Art Survey ofMotorTechnology Applicable to Hermetic Compressors for Domestic Refrigerators/Freezers. Draft report, prepared by ArthurD. Little, Inc., May, 1990. 4. Tecumseh Products Co., Tecumseh, MI, conference with J. Hackstedd, 1. 

### Hammond, T. Kandpal and HU Richardson at Tecumseh. 5. Robert Chilton, Danfoss, Inc., Mahwah, NJ, telcon 11/16/89. 6. Robert Hawley, VP Marketing Americold, White Consolidated Industries, Cullman, AL, telcon 11/22/89. 7. 1. 

### Vickerman, ManagerGE MotorMarketing, FortWayne, IN, telcon 12/7/89. 8. Philip Teague, ChiefHermetic Engineer, GE Small Motor Division, FortWayne, IN, conference at FortWayne, 1/31/90. 9. ADL report "Study ofEnergy-Saving Options for Refrigerators and Water Heaters," Vol. 1, May 1977, p. 86, adjusted perinputs from Tecumseh Products Co. (ref 1), 

### and W. Soedel, HerrickLaboratories, Purdue University. 10. M.C.C. Tsao, "Thermodynamics ofaHermetic Reciprocating Refrigeration Compressor with MCCT-Piston and Valve System," ASME Transactions, Vol. 103, April 1981. 11. Hideki Kawai, et aI, "The Development ofHigh Efficiency Compressors by Reducing Suction Gas Temperature," CompressorEngr. Division, Matsushita, Ltd., Fujisawa-shi Kanagawa, Japan. 12. GE Small MotorDivision, FortWayne, IN, conference with H. Harms and D. Erdman at FortWayne, IN, 1/31/90. 13. "Impact ofa CFC Ban on the Cost and Performance ofHousehold Refrigerators, Centrifugal Chillers, and CommerciallIndustrial Refrigeration Systems," Technical Memorandum to U.S. DOE by ArthurD. Little, Inc., August 1980. 14. Nelson, R.T. and MacCarthy, P.W., Research and DevelopmentofEnergy-Efficient Appliance Motor-Compressors, Final Report, Vol. I Executive Summary and Vol. 

### III Development and Field TestPlan. Columbus Products Corporation (division of White Consolidated Industries), Columbus, OH. 15. Prof. W. Soedel, conversation atHerrickLaboratories, Purdue University, 2/1/90, on compressorefficiencies and limits ofrolling piston and scroll compressors.

9-116. Prof. D. Tree, conversation at HerrickLaboratories, Purdue University, 2/1/90, general developments in the RlFcompressor field. 17. Laercio Hardt, Embraco, CalumentCity, IL office, telcon 5/10/90, compuressor efficiency status. 18. Prof. R. Hill, University ofMaine Orono, telcon 4/17/90, LaBreque cycle status, problems with R22 small compressor availability. 19. Constantio Netto, Embraco, Sao Paulo, Brazil, telcon, 3/9/90, status ofGE rotaries and problems encountered. 20. J. Sturgeon, GE CompressorDivision, Nashville, TN, telcon 3/9/90, availability of improved efficiency compressors. 21. E. Westhoff, White Consolidated Industries (Americold), telcon 3/2/90, advised that Americo1d does not have compressors with EERs in the 5 range for the lower capacities. 22. 

### 1. Poufve, GE Small Motors, FortWayne, IN, telcon 1/12/90, obtain data on current motorefficiencies. 23. R. Huffman, Consultant, telcon, 1/4/90, breakdown ofRlFmanufacturers and their compressor suppliers. 24. T. Nakashima, Sanyo sales office, Chicago, IL, telcon 12/28/89, solicitinformation on Sanyo compressors, current and future plans. 25. John McCullough, Arthur D. Little, Inc., telcon, 12/28/89, feasibility ofscroll compressors in RlFsizes, not optimistic. 26. Mark Okazaki, Matsushita (Panasonic), US rep, telcon 12/12/89, obtain information on Panasonic compressors, current and future plans. 27. 

### 1. Hackstedde, Tecumseh Products 

### Co., telcon 11/30/89. Tecumseh makes most motors, buys some from GE. 28. Donald Allen, PresidentAppliance Division, Emerson Motors, telcon 11/30/89, estimates RSIR motors in mid to high 70s, 5-6% more for capacitor run, Emerson no longer supplies to home RlFmarkets. 29. R. Chilton, Danfoss US MarketManager, tlecon, 11/16/89, supply to GE, Amana, White, Admiral, rotary costly to manufacture. Danfoss notconvinced rotary is right direction, pushing 134aas are Matsushita and Sanyo. 30. "Compressors," Ch. 12, 1983 edition ASHRAE Fundamentals.

9-231. 

### R Sauber, "Field Testing-Energy Saving Hermetic Compressors in Residential Refrigerators," ASHRAE Transactions, 1982. 32. "The Pressure's On," D. E. Simpson, November 1989, Appliance Magazine. 33. "16th International Congress ofRefrigeration," Paris 1983, preprints. 34. T. Shimizu, et aI, "Current Status and Trends in CompressorDevelopment," translation from Refrigeration, Vol. 62, No. 720, October 1987. 35. Laercio Hardt, Embraco, presentation atAHAM meeting, Battelle, Columbus, OH, 4/25 & 4/26/89. 36. D. Erdman, et aI, "Electronically Commutated Motors for the Appliance Industry," 1984 IEEE publication. 37. "RecentTechnological Trends and Lineup ofHitachi Compressors," K. Kikuchi, et aI, Hitachi Review, Vol. 36,1987. 38. Nagatomo, S., et al,"Performance Analysis ofRolling Piston Type Rotary Compressorfor Household Refrigerators," 1986 Purdue CompressorConference Proceedings, pp. 291-298. 39. van derWalt, N., Unger, R, SunpowerInc., Athens, Ohio, "The Simulation and Design ofa High Efficiency, LubricantFree, LinearCompressorfor a Domestic Refrigerator." 40. van derWalt, N., Unger, R., SunpowerInc., Athens, Ohio, "A High Efficiency, Oil Free, Linear Compressor, TestResults." 41. NASA Refrigeration Meeting. 42. Reefe, R, Americold, telecon 3/23/93, discuss test results of6.0 EER ECM motor compressor and ofSunpowerlinearfree piston compressor.

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