v>EPA
United States
Environmental Protection
Agency
Office of Atmospheric
and Indoor Air
Programs
EPA-430-R-93-010
June 1993
State of the Art Survey of
Hermetic Compressor
Technology Applicable to
Domestic
Refrigerator/Freezers
Recycled/Recyclable
TX /iO Printed on paper that contains
VH V at least 50% recycled fiber
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State Of The Art Survey
Of Hermetic Compressor Technology
Applicable to Domestic Refrigerator/Freezers
Prepared for
Environmental Protection Agency
Division of Global Change
Revised: March, 1993
Reference 64128
64122
67986
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Prologue
Compressor technology will play an important role in allowing potential efficiency
improvements of various refrigerator/freezer design alternatives. This review of the
technology of small hermetic refrigeration compressors was undertaken as part of a
larger study to evaluate the options for maximizing the efficiency of domestic
refrigerators. The compressor characteristics documented in the report are being used in
the overall study as part of the input database for modeling and evaluating design
options.
Findings
1. The efficiency of "large" compressors (capacity greater than 750 Btu/hr) has
been improved significantly over the past decade. Further improvements are
possible, at incrementally higher costs.
In 1980 the best "large" compressor energy efficiency ratio (EER), at standard
conditions, was 4.0 Btu/Watt-hr.
The best current production EER levels are 5.5.
In response to 1993 refrigerator/freezer efficiency 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 of commercially
available compressors in this capacity range. About one half of this increase is
attributable to higher motor efficiencies, the remainder to incremental reductions in
mechanical and thermal losses.
An EER level of approximately 6.5 is technically feasible, at an incremental increase
in OEM costs on the order of $ 15.
2. The value of these energy savings was calculated.
Assuming an electric energy cost of 80 per KwH, for real discount rates less than
10%, the real, present value of the saved energy is on the order of $100, over the
projected 15 year life of the refrigerator, much greater than the incremental cost, of
the increased efficiency compressor. '
3. Smaller compressors will be needed in future refrigerators.
Super insulation (either vacuum panels or thicker walls) is likely to reduce loads.
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Dual loop (or staged) systems using two compressors may replace single
compressor/evaporator designs to gain theoretical energy efficiency advantages.
4. Small compressors are inefficient because of a lack of economic incentive to
develop efficient models; substantial improvements can be obtained by
applying currently available technology.
Current small compressors are relatively inefficient, mainly due to inefficient motors
and high levels of suction gas heating.
Current technology 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. Energy improvements from improved small compressor performance are
possible with improved motors and application of other existing technology.
Dual loop and super-insulation systems should be economically viable with
this currently available (but as of yet) unmanufactured technology.
Small compressors can be improved 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/cost calculations
used in this report.
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Table 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 Other Regulatory and Policy Initiatives .................................. .................. 1-2
2.0 Issues in Compressor and Motor Selection .................. . ....... . ............... 2-1
2.1 Compressor Technology Options ..................................... ..... .................... 2-3
3.0 Compressor Technology: Current State of the Art . .......... . ............... ... 3-1
3.1 Description of Current Compressor Technology ............. ........................ 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 Compressor Costs [[[ - ........................ 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 Linear Free 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
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Table of Figures
Figure 2-1: Lorenz-Meutzner Refrigerator/Freezer Cycle 2-1
Figure 3-1: Typical Small Hermetic Reciprocating Motor-Compressor 3-2
Figure 3-2: Typical Small Hermetic Rotary Motor-Compressor 3-4
Figure 3-3: Compressor Efficiencies versus Capacity 3-10
Figure 3-4: Estimated OEM Price of Average Efficiency Refrigeration
Compressors 3-11
Figure 3-5: Ideal PV Processes at Standard Conditions 3-12
Figure 3-6: Conceptual Means of Isothermalizing the Compression
Process 3-13
Figure 3-7: R12 Pressure/Enthalpy Diagram Effect of Superheat 3-18
Figure 3-8: Compressor Heat Distribution 3-19
Figure 3-9: Low Side Compressor 3-20
Figure 3-10: Effect of Suction Gas Temperature on Compressor Work for
Constant Mechanical & Cylinder Efficiency 3-20
Figure 3-11: Low Side versus High Side Crankcase in a Single Cylinder
Compressor 3-22
Figure 4-1: Improvement in Efficiency of "Large" Compressors for
Domestic Refrigerator/Freezers over the Past Decade 4-1
Figure 4-2: Estimated Incremental Cost of Improved Efficiency Large
Compressors 4.4
Figure 5-1: Estimated Incremental Cost of Improved Efficiency Small
Compressors 5.4
IV
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Table of Tables
i
Table 2-1: Alternative Refrigerants 2-3
Table 3-1: Refrigerator/Freezer Manufacturers and Compressor Suppliers ... 3-8
Table 3-2: Partial List of Current High Efficiency Refrigerator/Freezer
Compressors ' 3-9
Table 3-3' EER Limiting Cases, Based on Alternate Compression Processes
..:. 3-13
Table 3-4: Typical Compressor Input Power Distribution ....: 3-14
Table 3-5: Motor Efficiencies r 3-15
Table 4-1: "Large" Reciprocating Compressor Input Power Distribution vs
Efficiency Level - 4-2
Table 4-2: "Large" Rotary Compressor Input Power Distribution vs
Efficiency Level 4-3
Table 5-1: Estimated Small Reciprocating Compressor Limiting Efficiency
and Input Power Distribution, Based on Loss Scaling 5-2
Table 5-2: Small (200 Btu/hr Nominal Capacity) Reciprocating Compressor
Input Power Distribution at Several Efficiency Levels 5-2
Table 5-3: Estimated Small Rotary Compressor Limiting Efficiency and
Input Power Distribution, Based on Loss Scaling 5-3
Table 5-4: Scaling of Losses in Reciprocating and Toary Compressor 5-5
Table 6-1: Energy Savings 6-1
Table 6-2: Present Value of Energy Savings Obtained with Increased
Efficiency Compressors , 6-3
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Acknowledgements
During the course of this work, a number of compressor, motor, and variable speed
drive manufacturers were contacted, whose assistance and cooperation we acknowledge
and appreciate:
Compressors: Tecumseh Products Co., Copeland, Americold, Danfoss, Embrace,
Sanyo, and Matsushita
Motors: General Electric, A. O. Smith, Emerson, and Baldor
Variable speed drives: Toshiba, Emerson, Mitsubishi, Hitachi, Westinghouse,
Magnetek, Lenze, Vee Arc, Ranco, Inland, PMI, Minarik, EG&G, Fasco, Boston
Gear, and Graham
vi
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1.0 Introduction
The design of the domestic refrigerator-freezer will be undergoing a significant set of
changes over the 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 need for a comprehensive examination of the
design options for domestic refrigerator/freezers. These interrelated options include
compressors, motors, refrigerants and lubricants, the refrigeration cycle, cabinet
insulation, and other aspects of cabinet thermal design.
This report covers the technology status of the small hermetic compressors used in the
refrigeration system of domestic refrigerators, addressing the potential for and cost of
improvements in the efficiency of small 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 that fully halogenated
chlorofluorocarbons (CFCs) have caused measurable deterioration of the atmosphere's
stratospheric ozone layer, which plays a significant role in attenuating solar ultraviolet
radiation. Increased levels of ultraviolet radiation would have a large number of
undesirable effects, including increased levels of skin cancers. Over the past few years,
this subject area has received renewed attention as the result of observations in the
mid-1980s of "gaps" in the ozone layer in the vicinity of the poles. As a result of this
attention, the Montreal CFC protocols were concluded in Fall, 1987. This international
agreement was signed and ratified by the major free world industrial nations requiring a
freeze, then phased production curtailments, of CFCs. By 1998, production of CFC-11,
CFC-12, CFC-113, CFC-114, and CFC-115, 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 of allowable levels of production accordingly. The
reassessment completed in 1990 resulted in a nearly total phase out of CFCs by the year
2000. The Clean Air Act of 1990 has codified this accelerated GFC phase out schedule
into U.S. environmental law. In April of 1991, NASA reported the results of
satellite-based measurements of stratospheric ozone levels indicating that ozone
depletion of 5% over the mid latitudes has already occurred. The result of this
development has been further acceleration of the timetable for CFC phase out. In
November, 1992, the Montreal Protocol Copenhagen Amendments accelerated the
complete phase out of CFCs to January 1, 1996. This has a direct impact on R/F
insulation and compressors which have been designed around the characteristics of
CFC-11 and CFC-12, respectively.
1-1
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1.2 Global Warming
Concurrent with the accumulation of scientific evidence of stratospheric ozone
depletion, increasing concerns have been developing about global warming caused by
increasing atmospheric concentrations of carbon dioxide and the trace greenhouse gases.
The most significant of the 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 or reduce CO2 emissions
have been proposed. Because CO2 is one of the basic combustion products of all of the
fossil fuels used to produce energy for heating, transportation, and electric power
generation, measures to reduce CO2 emissions require the burning of less 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 of consumer appliances (Federal Register, 1989a). Depending on the
category, the standards take effect between 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 most efficient products 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 of significantly reduced levels of allowable electric energy consumption for all
categories of domestic 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 took effect on 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 of total societal cost.
California, Oregon, Washington, most of the New England states, Wisconsin, and New
York have adopted ratemaking processes in which utility rates or return on investment is
adjusted so that utilities do not lose profits for forgone Kwh sales, but can profit from
demand reductions. In California and several other states, the non-pollution aspect of
demand reductions has led regulators to allow shared savings of customer bill reductions
to further increase utility profits. As a consequence of this change in utility regulation,
the demand for efficient refrigerators is rising.
1-2
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The Golden Carrot/early retirement program is one concrete manifestation of this trend.
Under the Golden Carrot, utilities are banding together to pool rebates to produce an
incentive for production of a R/F that is 30% better than DOE's 1993 standard in the 18
to 22 cubic foot range. With the impetus described earlier from CFCs, global warming,
and other state regulatory reforms, the Golden Carnot will provide a strong incentive for
vast improvements in energy efficiency.
1-3
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2.0 Issues in Compressor and Motor Selection
To meet the dual challenge of designing for new blowing agents and refrigerants, while
meeting much higher efficiency standards, significant changes in the design of both the
cabinet and the refrigeration cycle are under consideration. There are many design
options, and it is important to evaluate the potential of each option to contribute
significant energy savings, without using CFCs and without increasing the cost of the
refrigerator beyond the value added by the change (with respect to energy costs or
utility) or decreasing the utility of the refrigerator to the consumer. Design options that
are being pursued in current R&D programs include:
Low thermal loss cabinets ("super-insulated boxes") reduce energy consumption by
reducing the amount of cooling that is needed. To utilize heat exchangers effectively
and minimize cycling losses, the compressor capacity should be reduced in
proportion to the thermal load. However, sufficient compressor capacity 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 of the increased COP at the higher evaporator
temperature that can provide the required cooling of the fresh food compartment.
The Lorenz cycle, shown schematically in Figure 2-1 uses a rion-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 Refrigerator/Freezer Cycle
HX1
HX2
Condenser
i Fresh Food Evaporator r
Temp
Freezer Evaporator
Entropy
Variable speed compressor operation can save energy by allowing continuous
operation at low capacity, eliminating cycling losses and allowing more efficient
utilization of heat exchangers. Overspeed operation can provide additional capacity
for pulldown. The latter characteristic might be particularly advantageous with high
performance cabinets.
2-1
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One consequence of these design changes is a decrease in the compressor capacity
required to best match a given size refrigerator. Improved insulation and cabinet design
will reduce cooling loads, reducing required compressor capacity. Two compressor/two
evaporator systems meet the freezer and fresh food compartment heat loads with
separate refrigeration systems, obtaining, in theory, significant improvements in the
efficiency with which the load in the fresh food compartment is met (higher evaporating
temperatures, no defrost cycle). Even less compressor capacity is needed, especially in
the fresh food compartment where nominal compressor capacities of only 200 Btu/hr
may be needed. While the Lorenz cycle does not inherently result in a drastic
compressor capacity reduction, other compressor problems, 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/hr nominal
capacity) have very poor efficiencies, low enough in some cases to completely negate
the gains obtained from the design options described above. To realize the efficiency
benefits of reduced loads and dual evaporator systems will require improved efficiency,
lower capacity compressors.
Regardless of the R/F cabinet and refrigeration cycle design approach taken, increases in
the efficiency level that is available in refrigerant compressors will result in proportional
increases in the efficiency of the refrigerator using the compressor.
A major issue for compressor design is the change in refrigerant from CFC-12 to a low
ozone depletion, low global warming potential refrigerant. While CFC-12 has been
shown to be a significant part of the cause of both stratospheric ozone depletion and
global warming, it is an excellent working fluid for domestic refrigerator/freezers. It has
a favorable pressure-temperature relationship, good thermodynamic efficiency, stability,
total miscibility with low cost mineral oil, and moderate temperature rise with
compression and is non-flammable, non-toxic, and low cost. Alternate refrigerants that
do not contain chlorine or have currently acceptable ozone depletion potentials do not
possess identical attributes of CFC-12. Thus, compressor modifications or new designs
will be required to adapt to the characteristics of a selected alternative refrigerant. Table
2-1 lists some of the potential alternative refrigerants and their status, including
potential substitutes for CFC-11, CFC-114, and CFC-502, as well as for CFC-12.
The major working fluid options include the near drop in replacements for CFC-12 (i.e.,
those refrigerants having vapor pressure-temperature curves close to that of CFC-12),
lower vapor pressure refrigerants such as HCFC-124, and non-azeotropic refrigeration
mixtures (NARMs). Lower vapor pressure refrigerants might be utilized in low
capacity systems (design options described above), if shown to result in higher
efficiency of low capacity compressors, by virtue of the larger displacement that would
be needed. The Lorenz cycle would utilize a NARM.
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Table 2-1: Alternative Refrigerants
Substitute
Refrigerant
HCFC-22
HFC-134a
Ternary
HFC-152a
HFC-123
HFC-124
HFC-125
HCFC-141b
HCFC-142b
NH3
Displaced
Refrigerant
CFC-12
CFC-502
CFC-12
CFC-12
CFC-12
CFC-1 1
CFC-114
CFC-502
CFC-1 1
CFC-12, 114
CFC-1 1
CFC-12
CFC-502
Probable
Availability
Current
Current
(w/HCFC-124)
1993
Current
1990-1991
1993-95
Available in
limited amounts
Current
Current
Current
Description, Status, Comment
Commercially available, widely used refrigerant.
Contains chlorine, will be phased out under
Montreal Protocol Copenhagen Amendments
Commercially available, rapidly expanding
production
Available in limited amounts
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
Toxicity tests have shown sufficient toxicity to set
AEL at 10 ppm; commercially available
Co-product of HCFC-'1 23 production. Long term
toxicity testing started
Near-term availability in blends to replace CFC-502
Commercial production began in July 1988.
Toxicity testing underway. Possible use as a foam
blowing agent
Used in R22/R142b blends
Commercially available. Widely used in industrial
refrigeration sector. Toxic with low flammability
Source: Arthur D. Little, 1993 i
The major issues that need to be considered in adapting the compressor design to an
alternate refrigerant include 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 smaller capacities.
Motor technology is an important consideration, because increasing the efficiency of the
compressor motor is a straightforward way to improve compressor efficiency. For
variable speed compressors, the variable speed motor and electronic drive represent the
major technology component and the major cost driver.
2.1 Compressor Technology Survey
In view of the importance of compressor performance to the potential efficiency
improvements of various design alternatives, this review of the technology of small
hermetic refrigeration compressors was undertaken as part of a larger study to evaluate
2-3
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the options for maximizing the efficiency of domestic refrigerators. The compressor
characteristics documented in the report are being used in the overall study as part of the
input database for modeling and evaluating design options.
This report covers the technology status of small hermetic compressors used in the
refrigeration system of domestic refrigerators, addressing the potential for and cost of
improvements in the efficiency of small hermetic refrigeration compressors, with
particular emphasis on smaller (<800 Btu/hr nominal capacity) compressors.
This report is intended to serve three functions: 1) description of the state-of-the-art of
current compressors and the potential for future improvements; 2) summary of
performance and cost data as input to evaluations of refrigerator/freezer system design
options; and 3) present preliminary results indicating the level of R/F energy
consumption reductions that can be obtained through the use of high efficiency small
compressors.
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3.0 Compressor Technology: Current State of the Art
3.1 Description of Current Compressor Technology
Two types of compressors are used in current domestic RTF units, in the United States
and worldwide:
i
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 of oil and refrigerant
which could occur through rotating seals or mechanical fittings.
i
3.1.1 Reciprocating Compressors
The reciprocating compressor has been and still is the most common type used. It has
reached a high state of development, is mechanically efficient and reliable and is
relatively less expensive to manufacture than its alternative, the rotary, because of its
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 crankshaft which is an integral extension of the
driving motor shaft. The piston is connected to the crank by a connecting 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 of which, 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 of the motor shaft which supplies oil to
the rotating and reciprocating parts of the pump and motor. !
Gas is drawn from the R/F evaporator into the compressor shell in the space in the
cannister surrounding the motor/pump combination. It circulates, within the space aided
by the fan effect of the motor rotor. This cools the motor and the pump. A significant
portion of the heat picked 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.
Higher efficiency compressors may provide for a directed suction path from the inlet at
the shell directly into a muffler assembly, reducing the superheating of the suction gas.
Forced air cooling of the compressor shell may then be required to provide for motor
cooling. The path of the 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 shut down in the evaporator to be
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
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Figure 3-1: Typical Small Hermetic Reciprocating Motor-Compressor
Crankshaft
Discharge
Suction
Spring;
Oil Sump
Piston &
Connecting
Rod
Discharge
Valve & Port
Suction Valve
& Port
Electrical
Connector
Oil Pump
Inlet
Source: ADL sketch of Danfoss IL3A Compressor (315 Btu/hr, 3.5 EER)
3-2
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heat 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 a part of the pump
suction and discharge assemblies dampen gas pulsations.
3.1.2 Rotary Compressors
This class of compressors for R/F use in the range of interest (200 to 800 Btus) is
limited to the rolling piston - stationary vane type rotary compressor. Rotary vane
compressors were used in the past but are less efficient than either the reciprocating or
rolling piston compressors and are no longer employed.
The rolling piston compressor (Figure 3-2) consists of a cylindrical "piston" rolling on
the wall of a "cylinder" which is capped at both 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 of the motor shaft. A reciprocating 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 of the piston. A reed valve is used on
the discharge port to prevent back flow 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 R/Fs contain the motor and pump
housed in a welded shell hermetically 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 of the close
clearances, rotary compressors are very sensitive to ingestion of both particulates and
liquid slugs. To protect against both of these, all rotary compressors include a suction
accumulator and strainer close 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 of the heat dissipated in the motor is transferred to the
discharge gas. The motor stator is pressed into the shell which promotes conduction of
heat to the shell. Oil is 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 contact the cylinder walls or end caps, but is separated from
them by a thin oil film. Close clearances are necessary to prevent gas blow by. Sealing
of the clearances is enhanced by lubricating oil filling the clearance. Rotary
compressors generally have a suction accumulator/filter closely coupled to the inlet of
the pump to absorb liquid refrigerant and oil slugs at start up and prevent particulates
from entering the close clearances of the pump.
3-3
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Figure 3-2: Typical Small Hermetic Rotary Motor-Compressor
Rolling
Piston
Crankshaft
Eccentric
Oil Inlet
To Shaft
Port
Electrical
'Connector
Discharge
Motor Rotor
Motor Stator
& Windings
Discharge
Valve & Port
Suction Port
Oil Sump
Discharge
Port
Sliding Valve
and Spring
Suction Port
Source: ADL sketch of GE 2T/27RV/RS-36 Compressor (approximately 1000 Btu/hr)
3-4
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The 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 of the vane is not counterbalanced, but the product of the weight
and stroke length of the vane is at least an order of magnitude less than that of the piston
and connecting rod of a reciprocating compressor having the same capacity. The noise
level (vibration in the audible range) of a 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 of this pump is that it requires close clearances between
piston and cylinder, piston and crankshaft, crankshaft and support bearings and piston
and cylinder caps to prevent excessive gas blow by. These constraints require close
machining tolerances (on the order of tractions of a thousandth of an inch) over a much
greater total surface area than the reciprocating pump with attendant cost penalties and
make the pump less tolerant of wear.
There is a general expectation that the rolling piston compressor is not going to be as
efficient as a reciprocating pump in the lower capacity ranges. This results from the
inherently high surface to volume ratio of the cylinder and piston. This increases the
leakage paths and decreases the mechanical efficiency at a greater rate with reducing
size than it does for a reciprocating compressor. This opinion was supported by
Professor Soedel of the Herrick Laboratories at Purdue University (reference 15) and
extended to scroll compressors particularly because of the disproportionate increase in
leakage paths.
The principal advantages of the rolling piston pump are that it can be a high side pump
with attendant low suction gas superheat, high volumetric efficiency because of its
inherently low clearance volume and unrestricted suction port. These advantages in
energy efficiency are offset by higher levels of other losses, resulting in overall
efficiencies being less than the best current reciprocating pumps in the higher capacities.
Rotaries are more compact and lighter in weight than recips of comparable capacity; the
inherent potential cost savings have been a major reason for their growth in market
share during the 1980's.
'
3.1.3 Motor . .
Both types of hermetic compressors are driven by single phase squirrel cage induction
motors.
The motors consist of pairs of stationary (stator) electromagnets placed 180 degrees
apart, energized by "run" windings. The current produces an oscillating magnetic field
in the plane centered on the motor axis. This field induces a current in rotor conductors
which in turn generates a magnetic field in the rotor iron core. The rotor field opposes
the stator field producing a torque in the rotor.
3-5
-------
An induction motor depends upon the rotor magnetic field being generated by the
induced rotor current. This current is induced only when there is a speed difference
between the oscillating stator and rotating rotor fields.
The speed of the stator field is determined by the alternating current frequency and the
number of stator coil pairs; e.g., a 60 hertz current with a two coil (pole) stator produces
a magnetic field oscillating at 3600 ipm; a four pole motor at the same frequency
produces an 1800 rpm field speed. The motor speed is the speed of the rotor, which
must rotate at a speed less than that of the stator field. A speed difference, or "slip", of
approximately 3%, corresponds to the rated torque and power output of the motor. For
a two pole motor this speed is about 3500 rpm.
Both the stator and rotor magnets are made of laminated iron. The stator coils are
generally of wound copper wire. The rotor coil is formed of aluminum bars cast into
passages in the iron core resembling a squirrel 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 a balanced
force in the rotor with no net torque, consequently some assistance is required to start a
single phase motor. This is accomplished by a pair of electromagnetic 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 rotor has achieved a self sustaining
speed the start circuit is switched out of the 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 of the
circuit by a mechanical relay or a positive temperature coefficient resistor (PTCR). The
latter upon heating increases in resistance to reduce the current flow to negligible
proportions.
A motor with a starting winding only is called a split phase motor. A motor with a
resistor in series with the start winding is called a Resistor Start Induction Run (RSIR)
motor. These types produce a non-uniform rotating field which cause an objectionable
vibration in the rotor and must be switched out of the circuit after starting.
If a capacitor is 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 of motor is called a Capacitor Start (CS) motor.
A capacitor smaller than the starting capacitor can be inserted into the starting winding
circuit and the starting winding then operated continuously. This results in a higher
power factor which in turn reduces motor current, decreasing winding resistance losses
and improving motor efficiency by about 10%. This is called a capacitor run (CR)
3-6
-------
motor. If the motor also employs an additional capacitor in parallel with the "run"
capacitor for starting the combination is called a "Permanent Split Capacitor" (PSC)
motor. The "starting" capacitor is switched out of the circuit after starting.
The winding resistance, eddy currents and windage friction represent the losses of a
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), higher efficiency compressors for domestic R/Fs use
PSC motors. Smaller compressors generally use lower cost RSIR motors.
3.2 Performance
As discussed in Section 2, it is likely that compressors for future R/F designs will be
smaller than those current employed, approximately one half to one third of current
averages:
Present capacity range - 500 to 1500 Btu/hr
Probable future capacity range - 200 to 800 Btu/hr
A partial survey of the major U.S. refrigerator/freezer manufacturers and their
compressor suppliers (Table 3-1) shows that a wide selection of models in the 200 to
800 Btu/hr range is available to U.S. R/F 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." Current production state of the art represents a curve fit of the values of
compressors that are in full scale production and use by OEMs. Near term SOA is a
curve fit of the best EER 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: Btu/Watt-hr).
This expression is a variation of the "Coefficient of Performance" (COP) but well adapted to the units in
generalise 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, 90 F liquid and
suction vapor temperatures and 90° F ambient temperature.
3-7
-------
Table 3-1: Refrigerator/Freezer Manufacturers and Compressor Suppliers
R/F Manufacturers
General Electric/Hotpoint
Whirlpool
White/Frigidaire
Amana
Admiral/Maytag
Compressor Suppliers
General Electric
Danfoss
Tecumseh
Panasonic (Matsushita)
Embraco
Aspera
Americold
Panasonic
Sanyo
Tecumseh
Panasonic
Sanyo
Embraco
Tecumseh
Source: References 4, 24
3-8
-------
Table 3-2- Partial List of Current High Efficiency Refrigerator/Freezer Compressors
Manufacturer
Sanyo (CQ30)
Sanyo (C-M40L12C)
Tecumseh(AZ132OD)
Embraco (EMI 20ER)
Panasonic (S070LKAA)
Embraco (EM 30SC)
Panasonic (S090LKAA)
Embraco (EM 40SC)
Americold(ST104)
Embraco (EMI 40ER)
Panasonic (D112LRAA)
Panasonic (DA43L67)
Embraco (EM 55SC)
Embraco (EMI 55ER)
Americold(ST105)
Americold(HG106-1)
Panasonic (DA51L88R)
Panasonic (FN40R80R)
Americold(SI106)
Americold(HG107-1)
Tecumseh(AE1370W)
Panasonic (RA48L83R)
Panasonic (DA66L1 1 R)
Americold(ST107)
Americold(HG108-1)
Panasonic (RA53L11R)
Americold(HG 109-1)
Panasonic (DA73L13R)
Panasonic (FN60R12R)
Americold(HG110-1)
Panasonic (FN70R16R)
Americold(HG111-1)
Cap
139
180
200
220
278
315
349
420
440
440
476
516
565
580
580
621
635
651
660
708
740
754
806
810
848
850
947
973
1000
1029
1151
1193
Dspl
0.109
0.13
0.136
0.139
0.184
0.184
0.217
0.23
0.217
0.23
0.264
0.264
0.305
0.305
0.277
0.277
0.311
0.248
0.312
0.312
0.421
0.294
0.402
0.366
0.361
0.323
0.401
0.469
0.372
0.473
0.435
0.5
HER
2.32
2.6
3.13
3.6
3.1
3.89
3.5
4.3
4.7
4.4
4.4
4.75
4.41
4.5
5
5.35
5
4.46
5.06
5.4
4.6
4.86
5.15
5.06
5.5
4.8
5.53
5.25
4.61
5.51
4.79
5.55
Motor
RSIR
RSIR
RSIR
RSIR
RSIR
RSCF!
RSIR
RSIR
PTCR/CR
RSIR
RSCR
RSCR
RSCR
RSIR
PTCR/CR
PTCR/CR
RSCR
RSCR
PTCR/CR
PTCR/CR
PTCR/CR
RSCR
RSCR
PTCR/CR
PTCR/CR
RSCR
PTCR/CR
RSCR
RSCR
PTCR/CR
RSCR
PTCR/CR
Cooling
Static
Static
Static
Static
Static
Static
Static
Static
Fan
Static
Static
Static
Static
Static
Fan
Fan
Static
Static
Fan
Fan
Static
Fan
Fan
Fan
Fan
Fan
Fan
Fan
Fan
Fan
Fan
Fan
Type
Recip
Recip
Recip
Recip
Recip
Recip
Recip
Recip
Recip
Recip
Recip
Recip
Recip
Recip
Recip
Recip
Recip
Rotary
Recip
Recip
Recip
Rotary
Recip
Recip
Recip
Rotary
Recip
Recip
Rotary
Recip
Rotary
Recip
CAP Capacity (Btu/hr)
DSPL Displacement (cubic inches)
RECIP Reciprocating
RSIR Resistance Start Induction Run
PTCR/CR Positive Temperature Coefficient Resistor
Start/Capacitor Run
RSCR Resistance Start Capacitor Run
PSC Permanent Split Capacitor
Source: Engineering data supplied by the manufacturer of each model
3-9
-------
Figure 3-3: Compressor Efficiencies versus Capacity
5.5
4.5
ui 4
tu
3.5
2.5
CFC-12 Refrigerant
Current, State-of-the-Art
Large Recips -
Current, State-of-the-Art
Large Rotaries
Current, State-of-the-Art
"Small" Compressors
Static Cooled
Fan Cooled, Recip A
Fan Cooled, Rotary
_L
200
400 600 800
Cooling Capacity, Btu/hr
1000
1200
3.3 Compressor Costs
When sold on an OEM basis, refrigeration compressor prices are negotiated between the
compressor manufacturer and the refrigerator manufacturer. Compressor manufacturers
consider negotiated OEM pricing to be commercially/competitively sensitive
information, and are, therefore, reluctant to discuss the subject in 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 of OEM
price levels with several manufacturers.
3-10
-------
Figure 3-4: Estimated Price of Average Efficiency Refrigeration Compressors
40
30
20
10
200
400
600
Nominal Capacity, Btu/hr
800
1,000
1,200
3.4 Losses
As discussed above, the maximum efficiency level that is available in current,
production compressors, in terms of standard rating conditions EER, is 5.0 Btu/Watt-hr
for "large" (nominal capacity above 700 Btu/hr) and 3.5 for the small (200 Btu/hr)
compressors of specific interest to this study. The 5.0 EER of large compressors
represents a significant improvement over the efficiency levels that were available in
1980; much smaller improvements were made to the smaller compressors over this time
period. To place the following discussion of losses, and the discussion of potential
improvements (Sections 4 and 5) in perspective, it is instructive to consider the
thermodynamic limits on compressor EER.
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.
3-11
-------
Figure 3-5: Ideal PV Processes at Standard Conditions
250
200:
0.5
1 1.5 2
Specific Volume ft /Ibm
Curve A Curve B
2.5
Curve A: 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). Compressor efficiencies are
most commonly quoted relative to this thermodynamically "ideal" case. Implicit in the
selection of an adiabatic process as the standard is the presumption that heat rejection
directly from the compression process is impractical. For this process with CFC-12, the
ideal EER is 9.26, at 90°F suction temperature -10°F saturated suction temperature,
130°F saturated discharge temperature, and 90°F liquid and ambient air temperature.
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 of the compression process to this extent might
be approached by routing partially condensed refrigerant from the condenser to passages
in the cylinder body and head, then returning the refrigerant to the condenser, as shown
schematically in Figure 3-6.
The difference in performance between these two cases is relatively modest. While the
compression process of domestic refrigerator compressors is, in fact, not at all adiabatic,
with internal heat transfer and heat rejection from the compressor shell to ambient air
having a significant effect on the process, these heat transfer processes result in suction
gas heating, as well as isothermalization. For the purposes of the discussions in this
section and Sections 4 and 5, isentropic compression is taken as the "ideal" process.
Figure 3-5 was developed for CFC-12, the refrigerant that is the basis for all current
3-12
-------
Figure 3-6: Conceptual Means of Isothermalizing the Compression Process
Cooling
Suction
Gas
omp.l
in I
Compressor
/^=J"* (
-Jiff3
1 1 »-
Discharge
p .Jq
'3
i
-
Gas
Condenser
Liquid to capillary
tube/interchanger
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
fluids is small.
Table 3-3: EER Limiting Cases, Based on Alternate Compression Processes
Refrigerant
CFC-12
HFC-134a
HFC-152a
Cyclopropane
Ideal
Reversible Adiabatic
9.26
9.32
9.26
9.15
EER*
Isothermal 130T
9.8
9.8
9.8
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 compressor configurations, an approximate breakdown of
the electric input power among the basic categories of losses and useful (basis:
reversible adiabatic compression) refrigerant vapor compression. "The "high efficiency
reciprocating" compressor represents the highest efficiency compressors that are
currently available in production, in nominal capacities above 750 Btu/hr. Compressors
having this level of efficiency have been in commercial production for the past 2 or 3
years, and are the result of development efforts that were initiated in the late 1970s. The
"small, low efficiency reciprocating" compressor is representative of current
compressors whose nominal capacity is less than 400 Btu/hr. The "typical rotary" is
representative of both 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. It appears that by employing techniques that are now
3-13
-------
used in larger compressors, efficiencies of smaller compressors can be increased
significantly. Sections 4.0 and 5.0 discuss the potential for improvement for "large" and
"small" compressors, respectively.
Table 3-4: Typical Compressor Input Power Distribution
Input Electric Power
Motor Loss
Mechanical Losses
Suction Gas Heating
Discharge Port Losses
Low Side Pressure Losses
Compression and Expansion
Losses
Piston Blow By/Internal Leakage
Power to Compressor - Gas
Delivered"
High Efficiency
Reciprocating
(EER = 5.0)
100%
17%
8%
12%
2%
4%
2%
1%
54%
Small, Low
Efficiency
Reciprocating
(EER = 3.2)
1 00%
28%
10%
16%
2%
4%
4%
1%
35%
Typical Rotary
(EER = 4.7)
100%
22%
10%
3%
4%
2%
3%
5%
51%
3.4.7 Motors
Table 3-5 shows the efficiencies currently attained with compressor motors now in use
for the range of interest. It also shows a significant deterioration of efficiency that
parallels size. In this case, the reasons are largely due to less attention to performance in
smaller motors in the interest of low costs rather than an inherent trend to reduced
efficiency. GE, for example, says that it is not currently developing higher efficiency
motors for compressors below 600 Btu/hr because of lack of demand. Conversely,
because European and Japanese R/Fs are smaller than 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 efficient copper.
The possible efficiencies in Table 3-5 can be attained by the following improvements:
All motors operated as Permanent Split Capacitor motors,
Low eddy current and hysterisis loss steel through better chemistry and annealing,
Additional material (larger motors), i.e., more copper, bigger magnets, and
Different types of laminations in some motor designs.
Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet
conditions to discharge pressure.
3-14
-------
Table 3-5: Motor Efficiencies
Compressor
Rating
200 Btu
400 Btu
600 Btu
800 Btu
Motor
HP
1/16
1/8
1/6
1/4
Today's Motor
Type
RSIR
RSIR
RSIR&PSC
PSC
Efficiency
70-73%
73-76%
78-82%
80-84%
Possible
Type
PSC
PSC
PSC
PSC
Efficiency
84%
85%
86%
86%
RSIR
PSC
Legend
Resistance Start Induction Run
Permanent Split Capacitor
These improvements could result in motor costs from 1.5 to 3 times present costs, an
increase over current motor costs of approximately $10 to $25, according to General
Electric sources. The largest increases would be in the lower output motors used with
the smaller capacity compressors. These cost increases, while substantial, would still
leave the cost of smaller motors less than that of the current larger, more efficient sizes.
The projected efficiencies show an essentially level curve for the range of interest. This
effect alone would play a major role in flattening the compressor EER versus capacity
curve.
A separate report on the technology of both 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.
3.4.2.7 Mechanical Efficiency. Bearings comprise the major elements that determine
the mechanical efficiency of a pump. They consist of journal sliding and bearings.
Additional losses that might be categorized as mechanical losses include oil pumping
and windage, which together amount to 1 to 2 Watts.
Journal bearings are sleeve types and consist of main journal bearings on the crankshaft
for both types of pumps, connecting rod/journal and piston/journal bearings for
reciprocating and rotary pumps respectively and wrist pin/piston/connecting rod
bearings for reciprocating pumps. As a general rule, to minimize manufacturing costs,
refrigeration compressors have been designed as "members" of a "family" of
compressors covering a reasonably wide range of capacities, with as many common
parts as practical. Common parts include motor rotor and stator laminations, cylinder
housing castings, crankshafts, and connecting rod. A design that is mechanically
3-15
-------
optimized for the loads involved in the larger capacity models of the family will have
larger than optimum bearings, for example, in the smaller capacity models, and,
therefore, somewhat lower mechanical efficiency.
Sliding surface friction (with a lubricating oil film between parts) occurs between the
piston and cylinder of both reciprocating and rotary compressors. The sliding bearing
surface friction between the closely fitted cylinder and piston in a rotary pump is quite
high and accounts for the difference between the mechanical efficiencies of the two
types of pumps.
3.4.2.2 Pressure Losses. Low side pressure losses not only decrease the efficiency of
the compressor, but reduce the gas density in the cylinder at the end of the suction
stroke, reducing volumetric efficiency and cooling capacity. Major restrictions in the
inlet gas flow path of reciprocating compressors include the suction muffler and the
suction valve port/reed. The inlet to a rotary compressor is virtually unrestricted. The
net effect is to cause a slightly higher compression ratio than the external suction and
discharge conditions dictate.
Analogous pressure losses apply to the discharge side of the compressor, and result in
some overcompression.
The reciprocating pump has both a suction and discharge valve plate and smaller ports
while the rotary pump has only a discharge valve plate and large ports. These
distinctions account for some of the relative differences in efficiencies. Increasing port
sizes and valve plate sizes would benefit reciprocating pumps marginally. The value of
increased discharge valve port area in terms of reduced discharge pressure loss must be
balanced against the increased clearance volume of the discharge port. Rotary pumps,
however, have little room for improvement in this respect.
3.4.2.3 Volumetric Efficiency. Volumetric efficiency is the ratio of the 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 at the inlet to the compressor shell.
The major losses to volumetric efficiency are
Pressure losses between the inlet to the shell and the compressor cylinder (see brief
discussion above),
Suction gas superheating (see below). The combined effect of pressure losses and
suction gas superheating is to reduce the density of the refrigerant vapor in the
cylinder when the piston is at bottom dead center at the end of the suction stroke, and
Clearance gas reexpansion.
The clearance volume is comprised of a finite space between the piston and cylinder at
the end of the stroke and the discharge valve port. The piston clearance is necessary to
allow for the build up of manufacturing tolerances of the piston, cylinder housing,
crankshaft, and connecting rod as well as differential thermal expansion and insures
3-16
-------
against the 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 of the vane, on the discharge port side of the
vane.
i
The clearance volume retains a portion of the 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 of gas being drawn in from fully
occupying the cylinder. The effect of this phenomenon is to reduce the effective
capacity of the pump, a corollary of which is that the pump must ha.ve a larger
displacement than if its volumetric efficiency were 100%.
Volumetric efficiencies of reciprocating pumps tend to be in the 60% range, consistent
with the not insignificant magnitude in a reciprocating compressor of all three of the
above listed basic volumetric efficiency loss mechanisms. In contrast, the volumetric
efficiency of rotary compressors is typically over 90%, because the design minimizes
the basic losses, i.e., the direct suction minimizes both the heating of the 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 of the clearance gas is a reversible process. In reality, the
clearance gas compression and reexpansion processes are not reversible, primarily
because of cyclic heat transfer between the gas and the cylinder walls and head and
piston (the reexpanding clearance gas is somewhat cooler than during the compression
process, and returns less work to the piston than was originally required to compress the
clearance gas). Further, there is an indirect effect 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 of parts, the power input increased (because
less work was returned to the piston from reexpansion of the clearance gas), but the
refrigerant mass flow rate increased more in percentage terms, giving an overall
improvement in compressor EER.
For example, experimental work by Westinghouse reported in 1981 showed that a 760
Btu/hr compressor with a ratio of suction volume to clearance volume of 83 and an EER
of 3.4 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 of the gas at the inlet to the compressor shell,
the work of compression 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 of compression. (Note
3-17
-------
that both isentropic compression curves overstate the actual discharge temperatures that
would occur in a compressor having the indicated cylinder gas temperature, but is
generally illustrative of the effect on compression work of increased suction gas
superheat.)
Figure 3-7: R12 Pressure/Enthalpy Diagram Effect of Superheat - 90'F versus 230T
(-10T Saturated Suction, 130'F Saturated Discharge, Isentropic Compression)
aoo
200
408° F
140 ISO
Some superheat is desirable to prevent wet compression and for certain other
thermodynamic considerations; however, the suction gas is usually considerably
superheated as a result of ambient heating and subcooling of the liquid refrigerant for
increased refrigeration capacity. Figure 3-8 schematically shows the distribution of the
sources of suction gas superheat for reciprocating and rotary compressors.
In reciprocating compressors, the suction gas gains heat from a number of identifiable
sources. In a typical compressor, the suction gas enters the shell and mixes with gas that
has been heated by the motor and other higher temperature internal surfaces, primarily
the cylinder head and cylinder body and the discharge line. Next, the gas passes
through the suction muffler, which is often a chamber or pair of chambers bored into the
(high operating temperature) cylinder housing. Then the gas passes into the suction
3-18
-------
Figure 3-8: Compressor Heat Distribution
Reciprocating Compressor
hell cooling
Suction.
Gas
Q motor current
Rotary Compressor
motor current
Discharge
Gas
y II
Motor
Compressor
1 IZ
Qcomp A
JQ|n
Discharge
Gas
shell cooling
.Suction
Gas
manifold in the cylinder head, immediately adjacent to the high temperature discharge
manifold. Finally the gas passes into the cylinder, whose walls, have been heated to a
temperature intermediate to the inlet and discharge gas temperatures. As a result of
these multiple heating steps, the suction gas superheat may be as high as 150°F over the
gas temperature at the shell connection. The suction line can be directly connected to
the pump through the suction muffler provided that some type of leakage path permits
the pressure in the line to equalize with the shell gas. Figure 3-9 is an example of how
this could be accomplished. This is being done in some larger R/F compressors.
Matsushita reported a 6 to 10% improvement in compressor efficiency for a 25°F
reduction in superheat at the pump port by using this approach.
Figure 3-10 shows the theoretical increase of compressor work versus superheat over
the gas inlet temperature to the shell, for the case of a 90°F gas temperature to the
compressor shell. If cylinder gas temperatures were reduced from their typical 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 transfer paths
remain in the cylinder head 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 of the oil. To reduce suction gas
superheat further and lower pump temperatures, semi-hermetic designs could be
utilized. In semi-hermetic and open drive compressors, pump heat is rejected through
3-19
-------
Figure 3-9: Low Side Compressor
Joint Spring
Suction Tube
Suction Gas
Suction
Muffler
Compressor
Valve Head
Figure 3-10: Effect of Suction Gas Temperature on Compressor Work for Constant Mechanical &
Cylinder Efficiency
120 150 180 210 24O
Cylinder Suction Gas Temp Degree F at Beginning o1 Compression
270
3-20
-------
the 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 lower cylinder temperatures,
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 of the refrigerant charge in the system.
In rotary compressors, the suction gas passes from the compressor shell directly to the
cylinder inlet through a short (approximately 1 inch long) straight port in the cylinder
body (see Figure 3-4). Even with heat transfer to the cylinder and rolling piston
surfaces, suction gas superheat is very moderate in comparison with typical small
welded hermetic reciprocating compressors, usually on the order of 40° to 50°F. This
low level of superheat coupled with the negligible suction port pressure loss accounts
for the high volumetric efficiency that is typical of rotaries. This level of superheating
is so low that there is not much 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 compressor the crankcase or underside of the piston
operates at the suction pressure rather than the discharge pressure. The mechanical
effect of operating a single cylinder compressor crankcase/piston underside at high side
pressure (to allow direct suction to the compressor cylinder) is to significantly increase
the amount of work performed by the piston on each stroke (the theoretical network per
complete crankshaft resolution is the same in either case), 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 of the PV work is performed. With a high side crankcase, as shown
in the lower diagram, the refrigerant vapor in the cylinder on the top side of the 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 lower PV 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.
3-21
-------
Figure 3-11: Low Side versus High Side Crankcase in a Single Cylinder Compressor
P2
P1
P2
P1
Discharge
Gas
P1
Suction
Gas
* (during suction stroke)
Discharge
Gas
3-22
-------
4.0 Options for Improvement and Their Cost: Large Compressors
4.1 Reciprocating Compressors
The "large" reciprocating compressors currently used in the majority of U.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 of production
compressors has increased significantly over the past decade, as indicated in Figure 4-1.
In 1980, the maximum compressor EER in this capacity range was approximately 4.0
Btu/Watt-hr. By 1992, available EERs had increased from 5.3 to 5.5. Currently, in
early 1993, limited numbers of compressor samples having an EER of 6.0 (using a high
efficiency, electronically commutated permanent magnet DC motor) are available.
Figure 4-1: Improvement in Efficiency of "Large" Compressors for Domestic
Refrigerator/Freezers over the Past Decade (EER of Isentropic Compressor with 100%
efficient motor is 9.3 Btu/Watt-hr)
6
5.5
5
c-
I
at
1 4.5
S
a:
UI
4
3.5
3
1993 Samples
-
1990 Production SOA
1980 Production SI
. -
I I I
600 800 1000
Cooling Capacity (Btu/hr)
1400
Table 4-1 is a breakdown of the input power for large compressors at the four efficiency
levels shown in Figure 4-1. Significant reductions in motor losses and suction gas
heating losses provided most of the performance improvement between the 4.0 EER
level of 1980 and the 5.0 EER level of the best current production compressors. Further
improvement to the 5.3 EER level has come from incremental reductions of mechanical
losses (primarily through the use of a reduced viscosity lubricant), motor losses, and
suction gas heating. The OEM cost premium for this level of compressor efficiency is
on the order of $5. EER's on the order of 5.5 to 6.0 have been obtained through further
incremental loss reduction, as indicated in the last column of Table 4-1. The design
4-1
-------
measures include using the highest possible efficiency electric motor (electronically
commuted permanent magnet rotor DC motor) and reduced suction gas heating losses
(probably requiring active cooling of the cylinder body 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 of the compressor.
Table 4-1: "Large" Reciprocating Compressor Input Power Distribution vs. Efficiency Level
Motor Loss
Mechanical Losses
Suction Gas Heating
Discharge Port Losses
Low Side Pressure Losses
Compression and Expansion
Losses (clearance gas expansion)
Piston Blow By/Internal Leakage
Power Delivered to Gas
Compression
1980 Production
State of the Art
(EER = 4.0)
1 00%
23%
9%
15%
2%
4%
3%
1%
43%
Current
Production
(EER = 5.0)
100%
17%
8%
12%
2%
4%
2%
1%
54%
Current
Production
State of the Art
Samples
100%
16%
7%
11%
2%
4%
2%
1%
57%
Future Goa
(EER = 6.5)
1 00%
8%
6%
8%
2%
3%
2%
1%
70%
Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet
conditions to discharge pressure.
4.1.1 Non-Lubricated Linear Free Piston Reciprocating Compressor
In a project sponsored by the U.S. EPA, Sunpower has developed a prototype,
non-lubricated linear free piston compressor. The piston is driven by an
electronically-driven linear, permanent magnet motor. A measured EER at standard
conditions of 6.2 has been reported (References 40, 42). Efficiency improvements are
attributed to high (94%) motor efficiency, having only minimal friction losses in the
piston/cylinder gas bearing, offset to some extent by additional piston blow by and some
additional valve losses, both due to absence of the sealing effect of the oil film, piston
spring losses, and additional losses due to the unbalanced reciprocating motion of the
piston. The claimed potential for further efficiency improvement has not yet been
demonstrated. Significant issues include the level of reliability and operating life
(40,000 to 80,000 hours without failure in well over 90% of units is routine in current
compressors) that can be attained with an unlubricated free piston device and the cost to
manufacture, including necessary high energy permanent magnet materials and
electronic drive in the motor.
4.2 Rotary Compressors
In the "large" compressor capacity range, current production rotary compressors are
approximately competitive with current production reciprocating compressors with
4-2
-------
respect to both cost and efficiency; within the range of their own manufacturing
processes and cost/performance/reliability trade offs, different manufacturers have
placed greater or lesser emphasis on each technology.
Table 4-2 summarizes the breakdown of losses for current production and future,
improved efficiency "large" rotary compressors. The efficiency level of the middle
column was reported in Reference (38), and was the result of an intensive effort to
reduce all of the major losses (except the motor losses) to the minimum possible level.
A significant reduction in blow-by loss was achieved by reducing the rolling piston to
cylinder clearances by a factor of approximately two. Incremental reduction in
clearance gas re-expansion loss was obtained through reduction of the discharge port
volume and an incremental reduction of mechanical losses was achieved through
optimization of bearing diameters, lengths, and clearances. To date, no rotary
compressor is in production at this EER level, perhaps being indicative of the difficulty
in mass production of holding 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
Input Electric Power
Motor Loss
Mechanical Losses
Suction Gas Heating
Discharge Port Losses
Low Side Pressure Losses
Compression and Expansion
Losses
Piston Blow By/Internal Leakage
Power to Compressor - Gas
Delivered*
Typical Rotary
(EER = 4.7)
100%
22%
10%
3%
4%
2%
3%
5%
51%
Reference (38)
(EER = 5.1)
100%
22%
8%
3%
3 1/2%
2%
2 1/2%
4%
55%
With Best Motor
(EER = 5.6)
100%
16%
9%
3%
3 1/2%
2%
2 1/2%
4%
60%
* Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet
conditions to discharge pressure.
The last column 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 of the motor in a rotary compressor, the best motor
efficiency is approximately 2 percentage points less than that of the motor for a
comparable capacity reciprocating compressor. The resulting EER of 5.6 is essentially
comparable to the estimated limiting efficiency range for a "large" reciprocating
compressor.
4-3
-------
4.3 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 their customers. Figure 4-2 plots the estimated OEM
cost premium of increased efficiency for large compressors. The plot is based on
numerous informal discussions with compressor manufacturers and other industry
participants. A significant portion of the cost increase is attributable to incremental
increases in motor costs, as the maximum feasible efficiency of the motor is
approached.
Figure 4-2: Estimated Incremental Cost of Improved Efficiency Large Compressors
8
o
UJ
o
a
Q>
S
No attempt has been made to estimate the magnitude of the capital investments in R&D
and new tooling that would be required to develop new, higher efficiency compressors.
4-4
-------
5.0 Options for Improvement and Their Cost: Small Compressors
As indicated in the preceding sections, the efficiencies of the smaller compressors that
are the focus of this study and will be needed to effectively utilize improved cabinet
thermal design and dual evaporator cycles currently lag the efficiency that is available in
the best compressors utilized in larger refrigerators and freezers.: Where the efficiency
level of larger compressors has been improved substantially since 1980, only limited
improvements in efficiency have been implemented in the smallest compressors. In
order to achieve EERs in the 200 to 600 Btu capacity range comparable to those
currently available in the higher capacity 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 Btu/hr capacity range needed for dual loop systems.
Owing to the lack of current production, high efficiency, small compressors, a set of
loss scaling relationships have been developed to assess the extent to which the
efficiency improvements obtained in larger compressors might be obtained in small
compressors through similar design modifications.
5.1 Reciprocating Compressors
Table 5-1 is an estimate of the maximum efficiency level that is attainable in a small,
200 Btu/hr nominal 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 Btu/hr, EER 6.0) discussed in Section 4.1. The
resulting EER of 5.3 represents both a substantial improvement over current production
small compressors and a significant R&D challenge.
Consistent with the upper limits on small compressor performance estimated in the
preceding paragraph, Table 5-2 is a breakdown of the input power in a nominal 200
Btu/hr capacity compressor at current low efficiency levels, at an intermediate
efficiency level reached primarily through improved motor performance, 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.
5-1
-------
Table 5-1:
Estimated Small Reciprocating Compressor Limiting Efficiency and Input Power
Distribution, Based on Loss Scaling
Input Electric Power
Motor Loss
Mechanical Losses
Suction Gas Heating
Discharge Port Losses
Low Side Pressure Losses
Compression and Expansion
Losses
Piston Blow By/Internal Leakage
Power Delivered to Gas
Compression*
Large
Reciprocating
at Limiting
Efficiency
(EER = 6.5)
100%
8%
6%
8%
2%
3%
2%
1%
70%
Loss Scaling
Factor**
N/A
Table 3 - 5
Capacity "1/6
Capacity "1/s
--
--
Capacity "1fl
-
N/A
Loss Scaling
Factor @ 1/4
Capacity***
N/A
+2%
1.26
1.26
1.6
N/A
EER
Limiting
Efficiency @
200 Btu/hr
1 00%
10%
7.5%
10%
2%
3%
3.2%
1%
63%
«viw~iw MuiMh/Miiw wvi i ipi C7OOIUII Ul auiUCU UdlVC2fGU [llcloo HOW l9ie
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
Distribution at Several Efficiency Levels
i shell inlet
Power
Input Electric Power
Motor Loss
Mechanical Losses
Suction Gas Heating
Discharge Port Losses
Low Side Pressure
Losses
Compression and
Expansion Losses
Piston Blow By/Internal
Leakage
Power Delivered to Gas
Compression*
Low Efficiency, Current
Production SOA
(EER = 3.2)
100%
28%
10%
16%
2%
4%
4%
1%
35%
Intermediate Efficiency
(EER = 4.2)
100%
18%
10%
16%
2%
4%
4%
1%
45%
Maximum Efficiency
EER = 5.7
100%
10%
8%
10%
2%
3%
3%
1%
63%
Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet
conditions to discharge pressure.
5-2
-------
5.2 Rotary Compressors
Table 5-3 is an estimate of the maximum efficiency level that is attainable in a small,
200 Btu/hr nominal 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 Btu/hr, 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 of the extent to which scaling factors limit the efficiency
potential of rotary compressors for the smallest capacity applications.
Table 5-3: Estimated Small Rotary Compressor Limiting Efficiency and Input Power Distribution,
Based on Loss Scaling
Input Electric Power
Motor Loss
Mechanical Losses
Suction Gas Heating
Discharge Port Losses
Low Side Pressure Losses
Compression and Expansion
Losses
Piston Blow By/Internal Leakage
Power Delivered to Gas
"Large" Rotary
at Limiting
Efficiency
(800 Btu/hr,
EER = 5.6)
1 00%
16%
9%
3%
3 1/2%
2%
2 1/2%
4%
60%
Scaling
Factor**
N/A
Table 3 - 5
Capacity "1/6
Capacity ~1'3
-
--
Capacity '2/3
N/A
Loss Scaling
Factor @ 1/4
Capacity***
+2%
1.26
1.6
--
-
I
2.5
N/A
EEER
Small Rotary
Limiting
Efficiency @
200 Btu/hr
100%
18%
12%
5%
3 1/2%
2%
2 1/2%
10%
47%
4.4
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 their customers. Figure 5-1 plots the estimated OEM
cost premium of increased efficiency for small reciprocating compressors. The plot is
based on numerous informal discussions with compressor manufacturers and other
industry participants. A significant portion of the cost increase is attributable to
incremental increases in motor costs, as the maximum feasible efficiency of the motor is
approached.
No attempt has been made to estimate the magnitude of the capital investments in R&D
and new tooling that would be required to develop new, higher efficiency compressors.
5-3
-------
Figure 5-1: Estimated Incremental Cost of Improved Efficiency Small Compressors
5.4 Approach to Scaling of Losses
To estimate the potential efficiency level that could be obtained in small compressors
through the application of the same design features that have been used to improve large
compressor efficiencies, a set of simple scaling relationships relating the magnitude of
each loss to the capacity of the compressor was developed, for both reciprocating and
rotary compressors. Table 5-4 summarizes both sets of loss scaling relationships giving
the variation with capacity of the loss as a percentage of the total power input The
assumptions and rationale for the treatment of each loss are discussed briefly below
These scaling relationships generally apply to the 200 Btu/hr to 800 Btu/hr nominal
capacity range, at a constant 2-pole motor speed of approximately 3500 RPM A basic
assumption underlying these scaling relationships is that the small capacity design is
geometrically similar to the large capacity design, i.e. all dimensions are scaled by the
same factor, with the exception of close clearances. To a first order, the capacity and
input power are proportional to the displacement, which is proportional to the cube of
the linear dimensions, e.g. diameter, length. This simplified treatment ignores many
complexities and interrelations of important design variables, such as the temperature
dependence of oil viscosity; as such it can be regarded in the aggregate as being
generally indicative of the magnitude of the effect of compressor capacity on the losses
and the resulting efficiency potential.
5-4
-------
Table 5-4: Scaling of Losses in Reciprocating and Rotary Compressor
Motor Loss
Mechanical Losses
Suction Gas Heating
Discharge Port Losses
Low Side Pressure Losses
Clearance Gas Reexpansion
Piston Blow By/Internal Leakage
Scaling Rule for Loss vs. Capacity
Reciprocating
Small - Table 3-5
Capacity '1/6
Capacity'1'6
No Scale Effect
No Scale Effect
Capacity '1/a
No Scale Effect
Rotary
Capacity '1/e
Capacity "1/3
No Scale Effect
No Scale Effect
No Scale Effect
Capacity 'M
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 a rotary compressor is about 2 percentage points
less than the values indicated in Table 3-5, because of the higher operating temperature
and larger aspect ratio of the motor.
High and Low Side Pressure Losses. For both recips and rotaries, the effect of
geometrically scaling, at constant rotational speed, vapor passages and valve port
diameters is to reduce gas velocities and pressure losses, complicated by considerations
such as the effect of inlet and discharge valve reed stiffness on pressure drops. The
overall effect of scale 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, smaller mechanical losses include oil pumping and windage.
Piston friction power dissipation is the product of the average oil film shear force
acting on the piston (oil viscosity x piston sliding surface area x average velocity -
clearance) and the average piston velocity. The basic scaling relationships are:
- Average velocity ^ diameter (
- Piston sliding surface area <*= diameter squared
- Oil viscosity is constant
- The diametral piston to wall clearance is constant, on the order of 0.0001 to
0.0002 inches, dictated by the minimum practical level for selective assembly
- Overall, the power dissipation is proportional to d4, the ratio of power dissipated
to input power is d4/d3 = d= capacity173
5-5
-------
Shaft bearing friction: for a given journal bearing lubricant viscosity, L/D ratio,
clearance to diameter ratio, and operating eccentricity, at constant rotational speed
the power dissipated in the bearing is proportional to the design load, which varies
with the cube of the shaft diameter. The load, which originates primarily from gas
pressure acting on the piston is proportional to the piston diameter squared. The
ratio of the power loss (proportional to the bearing loads) to the total power input is
d2/d3 = d'1 = capacity-10.
The net effect of the relatively small piston friction plus the bearing friction is
estimated to be: mechanical power losses/total input power = capacity"1'6.
Suction Gas Heating. As discussed below, under rotaries, suction gas heating varies
with surface area to volume ratio, given a constant internal temperature distribution. In
small recips, suction gas heating is a more significant loss, and affects the temperature
distribution. With forced air cooling of the compressor shell, average internal
temperatures will be reduced to a greater extent at smaller capacities. Overall, a lesser
dependence of suction gas heating on scale is assumed: capacity"1'6.
Clearance Gas Reexpansion. The magnitude of the ratio of this loss to total input
power depends on the ratio of the clearance volume to the displacement. In larger, high
efficiency compressors, this ratio has been reduced to the minimum practical level', and
involves the handling of piston, cylinder, and connecting rod length tolerances, with a
resulting minimum piston crown to cylinder head clearance at top dead center. The
minimum tolerance based clearance will not scale down with reduced dimensions to any
significant extent. The piston stroke does scale, so the ratio of clearance volume to
displacement will be proportional to d"1 or capacity to "1/3.
Piston Blow-by. In a recip, the clearance between the piston and cylinder is 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.
5-6
-------
5.4.2 Rotary Compressors
Mechanical Losses. Crankshaft bearing loads and losses and piston sliding losses
scale in the same fashion as with recips. (overall loss to power output ratio scales with
capacity"1'6).
Suction Gas Heating. This is a fairly small loss in a rotary, and consequently
moderate changes in the loss will not have a significant effect om temperature levels
throughout the compressor. The magnitude of the heat transfer from 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 d2/d3 = d"1 = capacity"1'3.
Clearance Gas Reexpansion. The clearance volume in a rotary will scale with the
overall dimensions; therefore no significant scale effect is expected.
Piston Blow-By. This is a major loss in "large" rotaries, kept under control by rigorous
control of all internal clearances to the minimum practical level. It is doubtful whether
significantly smaller clearances can be maintained in smaller capacity rotaries. Based
on the assumption that the clearances are constant, the blow by loss is proportional^o
the perimeter which is proportional to d. The ratio of the loss to the capacity is d/d =
*5 ?/^
d = capacity .
5-7
-------
-------
6.0 Potential Value of Improvements
Table 6-1 summarizes the annual energy savings obtainable through improved
compressor efficiencies. These savings have been evaluated for two basic R/F (top
mounted freezer) configurations: single evaporator/single compressor and dual loop.
The baseline, single evaporator configuration is representative of a configuration 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
4.85"
5.0
5.5
6.0
3.6/4.7
4.6/5.1
5.3/5.5
Comment
Readily available
Best production
compressor
1 993 production
Approaching practical
limits
Currently available
With efficient motor
Approaching practical
limits
Annual Energy
kWh*
685
669
(518
570
(302
I537
498
Annual Energy
Saving** kWh
16
67
115
83
148
187
At DOE test conditions, taken to be representative of actual use
Annual energy savings relative to large compressor/single evaporator baseline
Cabinet volumes (cubic feet):
- freezer 4.6
- fresh food 13.4
- total 18.0
- adjusted volume 20.9
I
Federal energy efficiency standard:
- 1990 962kWh/yr
- 1993 690kWh/yr
4.85 EER, 893 Btu/hr compressor (readily available, moderate cost, reasonable
efficiency, provides typical pulldown capacity).
Cabinet insulation is polyurethane foam with R=8°F/in per Btu/hr-ft2
- average freezer wall thickness: 2 3/8 inch
- average fresh food compartment wall thickness: 1 3/4 inch
- insulation thickness in doors: 1 5/8 inch
Standard evaporator (2 rows deep x 8 rows high)
6-1
-------
Efficient (PSC type) fans:
- evaporator 50 CFM, 6.8 Watts
- condenser 90 CFM, 6.8 Watts
Electric resistance anti-sweat heaters: 5.5 Watts for mullion (the energy analysis
assumes that the heater power is on one half of the time to approximate the DOE test
procedure)
Vapor-line anti-sweat heat for cabinet door flange (cycle average of 4 Watts)
Defrost energy is assumed to be constant across all cases studied. Variations in cycle
time are assumed to be compensated by adjustments to the compressor timer interval.
This baseline configuration is similar to the baseline configuration that was used in the
Motor Technology Study (Reference 3) to evaluate efficiency improvements obtainable
from variable speed compressor and fan operation. The dual loop configuration uses
the same cabinet as the single evaporator configuration. The differences are:
Evaporators - each evaporator is the same size and capacity of the single evaporator.
Fans - each evaporator and condenser fan 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
For each of these basic configurations, the annual energy consumption at the DOE test
conditions was calculated for the baseline compressor efficiency, and for progressively
increased compressor efficiency, as indicated in Table 6-1, consistent with the
discussions in Sections 4 and 5. For both single evaporator and dual loop systems, the
annual energy savings are relative to the baseline single evaporator configuration.'
The value of these energy savings is tabulated in Table 6-2 assuming an electric energy
cost of 80 per kWh, for a range of real (after inflation) discount rates between 2 and 10
percent, over the projected 15 year life of the refrigerator. The discount rates cover a
range between real, after tax returns to savings accounts at the low end of the range, to
after inflation credit card interest rates at the high end of the range. For real discount
rates in the range that consumers would rationally choose for safe investments (at the
low end of the range), the present value of the saved energy is on the order of $50 to
$100 greater than the incremental cost, at retail, of the increased compressor efficiency.
6-2
-------
Table 6-2: Present Value of Energy Savings Obtained with Increased Efficiency Compressors
Large Compressor
Single Evaporator
Small Compressors
Dual Loop
Compressor
EER
4.85*
5.0
5 5
6.0
3.6/4.7
4.6/5.1
5.3/5.5
Savings
kWh/yr"
16
67
115
83
148
187
Present Value for Discounlt
Rate
2%
17
70
120
87
155
196
5%
14
58
99
71
127
161
10%
10
43
73
3
94
169
Cost of
Improvements
(in U.S. Dollars
at Retail)
0
5
10
20
20
40
80
Baseline
Relative to baseline
6-3
-------
-------
7.0 Areas f or R&D
As discussed in the preceding sections, considerable potential exists for improving the
efficiency of the small hermetic compressors that would be used in. dual loop designs.
The areas for improvement include increasing motor efficiency 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 major issues involved from the industry viewpoint are
product engineering and marketing issues of cost vs. market size, and integration with
the compressor Beyond these commercialization issues, provision of a high etticiency
motor for a small compressor is relatively straightforward. Because the motor is an
integral part of the compressor package, high efficiency motor development is best
carried out in the context of high efficiency compressor development; integration of a
high efficiency motor with an existing small compressor represents a low risk, readily
implementable means of obtaining a significant increase in small compressor efficiency.
Early development of small, increased efficiency compressors would provide important
support to current programs to demonstrate the performance potential of R/F s with dual
loop systems and/or super insulated cabinets. The focus of a project should be the
small 200 Btu/hr nominal capacity (at standard rating conditions) compressor intended
for the fresh food compartment of a dual loop refrigerator/freezer of typical (18 to 22
cubic feet) refrigerated volume.
The results of the performance and payback analysis discussed in Section 6 clearly
indicate that significant energy savings can be obtained with a dual loop refrigeration
system in a refrigerator/freezer, if the compressor efficiencies are the levels that can be
attained with the improvements discussed in Sections 4 and 5, with the energy saving s
at current electricity prices providing a 3 to 5 year payback of the projected $40 to $80
increase (at retail) in compressor cost. To support ongoing efforts to develop and
demonstrate this performance, early development of a small, maximum efficiency
compressor is justified. 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 smaU compressor designs to significantly reduce suction gas
superheating. Other loss areas should also be systematically examined tor improvement
potential. With the continuing acceleration of the CFC phase out, a non-CFC refrigerant
and compatible lubricant should be utilized.
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8.0 Conclusions
This study has dealt primarily with the extent to which the efficiency of small 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 of the art (in production) for larger capacity welded hermetic domestic
refrigeration compressors.
The efficiency levels now attained in the compressors used in the majority of domestic
refrigeration freezer applications (EERs between 4.8 and 5.3 Btu/Wat.t-hr, for nominal
capacities between 700 and 1100 Btu/hr) are the result of development efforts that were
initiated in the late 1970s. The hermetic refrigeration compressors in this capacity range
that were available then had EER values similar to today's small compressors.
Development efforts have been focussed on achieving significant efficiency gains, but
within very tight cost constraints and high reliability requirements. Mass scale
production and specification in R/Fs of the resulting high efficiency (but somewhat
higher cost) compressors has been driven by a combination of consumer demand for
higher efficiency refrigerators and the Federal energy efficiency standards.
In response to continuing tightening of the energy efficiency standards, compressor
manufacturers have continued to invest in the development of higher efficiency, large
compressors. Preproduction samples having an EER of 5.3 to 5.5 became available to
R/F manufacturers in 1991. Production of these compressors began in 1992. In
anticipation of tighter standards in the future, there are continuing R&D programs aimed
at developing still higher efficiency compressors. Samples of ECM motor driven
compressors having an EER of 6.0 have become available in early 1993. The practical,
physical limit for the EER of compressors 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 of these small refrigerators that is comparable to
market cost pressures (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 of the compressor is a significant increment in cost).
Efficiency in the 200 to 600 Btu range of compressors can be improved with the
application of current technology. Reciprocating compressors can be scaled down to
low capacity levels with much less efficiency penalty than rotaries. Efficiency levels on
the order of 5.0 EER can be attained at the small end of this range through application of
current technology. The major design changes needed to reach this performance level
are:
Substitution of higher efficiency PSC or RSCR motors for the RSIR motors used in
current small compressors,
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Implementing design features to reduce suction gas superheating in the compressor
including direct suction through an insulated muffler and forced air cooling of the '
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 smaller capacities (instead of the larger compressors in the
"family" of compressors of which the small compressor is a member).
The OEM cost premium of an improved efficiency, 200 Btu/hr nominal capacity
compressor would be on the order of $10 to $20 over the typical current OEM cost of
approximately $20 to $25. Regardless of economic payback time, these improvements
in small compressor efficiency may be a necessity for meeting the 1993 Federal energy
efficiency standard for domestic refrigerators and freezers, or more stringent future
standards.
Two generic types of compressors are now in common use in welded hermetic
compressors use in domestic refrigerators, reciprocating and rotary. For the smallest
required capacities, e.g., for the fresh food compartment of a dual evaporator system,
reciprocating compressors are the more promising option for attaining high efficiencies.
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9.0 References and Sources
1. Federal Register, April 21, 1989a, pp. 6077-8.
I
2. Federal Register, November 17, 1989b, pp. 47916-47938.
3. Dieckmann, J. and McMahon, E., State of the Art Smvp.v nf Motor Technology
Applicable to Hermetic Compressors for Domestic Refrigerators/Freezers. Draft
report, prepared by Arthur D. Little, Inc., May, 1990.
4. Tecumseh Products Co., Tecumseh, MI, conference with J. Hackstedd, J.
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. J. Vickerman, Manager GE Motor Marketing, Fort Wayne, IN, telcon 12/7/89.
8. Philip Teague, Chief Hermetic Engineer, GE Small Motor Division, Fort Wayne,
IN, conference at Fort Wayne, 1/31/90.
9. ADL report"Study of Energy-Saving Options for Refrigerators and Water Heaters,"
Vol. 1, May 1977, p. 86, adjusted per inputs from Tecumseh Products Co. (ref 1),
and W. Soedel, Herrick Laboratories, Purdue University.
10. M.C.C. Tsao, "Thermodynamics of a Hermetic Reciprocating Refrigeration
Compressor with MCCT-Piston and Valve System," ASME Transactions, Vol. 103,
April 1981.
11. Hideki Kawai, et al, "The Development of High Efficiency Compressors by
Reducing Suction Gas Temperature," Compressor Engr. Division, Matsushita, Ltd.,
Fujisawa-shi Kanagawa, Japan.
12. GE Small Motor Division, Fort Wayne, IN, conference with H. Harms and D.
Erdman at Fort Wayne, IN, 1/31/90.
13. "Impact of a CFC Ban on the Cost and Performance of Household Refrigerators,
Centrifugal Chillers, and Commercial/Industrial Refrigeration Systems," Technical
Memorandum to U.S. DOE by Arthur D. Little, Inc., August 1980.
i
14. Nelson, R.T. and MacCarthy, P.W., Research anH Development, of Energy-Efficient
Appliance Mntnr-Compressors. Final Report, Vol. I Executive 'summary and Vol.
Ill Development and Field Test Plan. Columbus Products Corporation (division of
White Consolidated Industries), Columbus, OH. |
15. Prof. W. Soedel, conversation at Herrick Laboratories, Purdue University, 2/1/90,
on compressor efficiencies and limits of rolling piston and scroll compressors.
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16. Prof. D. Tree, conversation at Herrick Laboratories, Purdue University, 2/1/90,
general developments in the R/F compressor field.
17. Laercio Hardt, Embrace, Calument City, IL office, telcon 5/10/90, compuressor
efficiency status.
18. Prof. R. Hill, University of Maine 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 of GE rotaries
and problems encountered.
20. J. Sturgeon, GE Compressor Division, Nashville, TN, telcon 3/9/90, availability of
improved efficiency compressors.
21. E. Westhoff, White Consolidated Industries (Americold), telcon 3/2/90, advised that
Americold does not have compressors with EERs in the 5 range for the lower
capacities.
22. J. Poufve, GE Small Motors, Fort Wayne, IN, telcon 1/12/90, obtain data on current
motor efficiencies.
23. R. Huffman, Consultant, telcon, 1/4/90, breakdown of R/F manufacturers and their
compressor suppliers.
24. T. Nakashima, Sanyo sales office, Chicago, IL, telcon 12/28/89, solicit information
on Sanyo compressors, current and future plans.
25. John McCullough, Arthur D. Little, Inc., telcon, 12/28/89, feasibility of scroll
compressors in R/F sizes, not optimistic.
26. Mark Okazaki, Matsushita (Panasonic), US rep, telcon 12/12/89, obtain information
on Panasonic compressors, current and future plans.
27. J. Hackstedde, Tecumseh Products Co., telcon 11/30/89. Tecumseh makes most
motors, buys some from GE.
28. Donald Allen, President Appliance 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 R/F markets.
29. R. Chilton, Danfoss US Market Manager, tlecon, 11/16/89, supply to GE, Amana,
White, Admiral, rotary costly to manufacture. Danfoss not convinced rotary is right
direction, pushing 134a as are Matsushita and Sanyo.
30. "Compressors," Ch. 12, 1983 edition ASHRAE Fundamentals.
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31. 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 of Refrigeration," Paris 1983, preprints.
34. T. Shimizu, et al, "Current Status and Trends in Compressor Development,"
translation from Refrigeration, Vol. 62, No. 720, October 1987.
35. Laercio Hardt, Embraco, presentation at AHAM meeting, Battelle, Columbus OH
4/25 & 4/26/89.
36. D. Erdman, et al, "Electronically Commutated Motors for the Appliance Industry "
1984 IEEE publication.
37. "Recent Technological Trends and Lineup of Hitachi Compressors," K. Kikuchi, et
al, Hitachi Review, Vol. 36,1987. "
38. Nagatomo, S., et al,"Performance Analysis of Rolling Piston Type Rotary
Compressor for Household Refrigerators," 1986 Purdue Compressor Conference
Proceedings, pp. 291-298.
39. van der Walt, N., Unger, R., Sunpower Inc., Athens, Ohio, "The Simulation and
Design of a High Efficiency, Lubricant Free, Linear Compressor for a Domestic
Refrigerator."
I
40. van der Walt, N., Unger, R., Sunpower Inc., Athens, Ohio, "A High Efficiency, Oil
Free, Linear Compressor, Test Results."
41. NASA Refrigeration Meeting.
42. Reefe, R., Americold, telecon 3/23/93, discuss test results of 6.0 EER ECM motor
compressor and of Sunpower linear free piston compressor.
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