United States

Environmental Protection

Agency

Office of Atmospheric

and Indoor Air

Programs

EPA-430-R-93-010

June 1993

ft

0 tate t rvey f

Herm iiiI r r

Tech ology p licable to IIIl

C

R frigerato I re zer

ffi RecycledIRecyclable n]:) Printed on paper that contains DV at least 50% recycled fiber

 CE-RM-93-801 COMMENT OR4
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

 CE-RM-93-801 COMMENT OR4
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 BtuIWatt-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 8¢ 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.

 CE-RM-93-801 COMMENT OR4
" 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.

ii

 CE-RM-93-801 COMMENT OR4
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

5.2 Rotary Compressors 5-3

5.3 Estimated Cost of Improved Efficiency Small Compressors 5-3

5.4 Approach to Scaling of Losses 5-4

5.4.1 Reciprocating Compressors 5-5

5.4.2 Rotary Compressors 5-7

6.0 Potential Value of Improvements

7.0 Areas for R&D

8.0 Conclusions

9.0 References and Sources

6-1

7-1

8-1

9-1

iii

 CE-RM-93-801 COMMENT OR4
Figure 2-1:

Figure 3-1:

Figure 3-2:

Figure 3-3:

Figure 3-4:

Figure 3-5:

Figure 3-6:

Figure 3-7:

Figure 3-8:

Figure 3-9:

Figure 3-10:

Figure 3-11:

Figure 4-1:

Figure 4-2:

Figure 5-1:

Table of Figures

Lorenz-Meutzner Refrigerator/Freezer Cycle 2-1

Typical Small Hermetic Reciprocating Motor-Compressor 3-2

Typical Small Hermetic Rotary Motor-Compressor 3-4

Compressor Efficiencies versus Capacity..................................... 3-10

Estimated OEM Price of Average Efficiency Refrigeration

Compressors 3-11

Ideal PV Processes at Standard Conditions 3-12

Conceptual Means of Isothermalizing the Compression

Process 3-13

R12 Pressure/Enthalpy Diagram Effect of Superheat 3-18

Compressor Heat Distribution 3-19

Low Side Compressor 3-20

Effect of Suction Gas Temperature on Compressor Work for

Constant Mechanical & Cylinder Efficiency 3-20

Low Side versus High Side Crankcase in a Single Cylinder

Compressor 3-22

Improvement in Efficiency of "Large" Compressors for

Domestic Refrigerator/Freezers over the Past Decade 4-1

Estimated Incremental Cost of Improved Efficiency Large

Compressors 4-4

Estimated Incremental Cost of Improved Efficiency Small

Compressors 5-4

iv

 CE-RM-93-801 COMMENT OR4
Table 2-1:

Table 3-1:

Table 3-2:

Table 3-3:

Table 3-4:

Table 3-5:

Table 4-1:

Table 4-2:

Table 5-1:

Table 5-2:

Table 5-3:

Table 5-4:

Table 6-1:

Table 6-2:

Table of Tables

Alternative Refrigerants 2-3

Refrigerator/Freezer Manufacturers and Compressor Suppliers 3-8

Partial List of Current High Efficiency Refrigerator/Freezer

Compressors 3-9

EER Limiting Cases, Based on Alternate Compression Processes

........................................................................................................ 3-13

Typical Compressor Input Power Distribution 3-14

Motor Efficiencies 3-15

"Large" Reciprocating Compressor Input Power Distribution vs

Efficiency Level 4-2

"Large" Rotary Compressor Input Power Distribution vs

Efficiency Level 4-3

Estimated Small Reciprocating Compressor Limiting Efficiency

and Input Power Distribution, Based on Loss Scaling 5-2

Small (200 Btu/hr Nominal Capacity) Reciprocating Compressor

Input Power Distribution at Several Efficiency Levels 5-2

Estimated Small Rotary Compressor Limiting Efficiency and

Input Power Distribution, Based on Loss Scaling 5-3

Scaling of Losses in Reciprocating and Toary Compressor 5-5

Energy Savings 6-1

Present Value of Energy Savings Obtained with Increased

Efficiency Compressors 6-3

v

 CE-RM-93-801 COMMENT OR4
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, Embraco,

Sanyo, and Matsushita

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

• Variable speed drives: Toshiba, Emerson, Mitsubishi, Hitachi, Westinghouse,

Magnetek, Lenze, Vee Are, Ranco, Inland, PMI, Minarik, EG&G, Fasco, Boston

Gear, and Graham

vi

 CE-RM-93-801 COMMENT OR4
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-1l3, CFC-1l4, and CFC-1l5, as well as certain "halons" were to be

reduced to 50% of 1986 levels. An additional provision provided for periodic review of

scientific evidence and adjustment 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 CFC 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 R1F

insulation and compressors which have been designed around the characteristics of

CFC-ll and CFC-12, respectively.

1-1

 CE-RM-93-801 COMMENT OR4
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

 CE-RM-93-801 COMMENT OR4
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 RlF 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

 CE-RM-93-801 COMMENT OR4
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 non-azeotropic

refrigerant mixture with two evaporators and an interchanger to operate the fresh

food compartment and the freezer at separate evaporator temperatures for higher

efficiency. (Lorenz, 1975)

Figure 2·1: Lorenz-Meutzner RefrigeratorlFreezer Cycle

Temp

Condenser 2

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

HX1

HX2

5 Freezer Evaporator Entropy

.. Variable speed compressor operation can save energy by allowing continuous

operation 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

 CE-RM-93-801 COMMENT OR4
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 tluid for domestic refrigerator/freezers. It has

a favorable pressure-temperature relationship, good thermodynamic efficiency, stability,

total miscibility with low ccst mineral oil, and moderate temperature rise with

compression and is non-tlammable, non-toxic, and low cost. Alternate refrigerants that

do 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-ll, CFC-114, and CFC-502, as well as for CFC-12.

The major working tluid 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.

2-2

 CE-RM-93-801 COMMENT OR4
Table 2-1: Alternative Refrigerants

Displaced Probable Description, Status, Comment

Refrigerant Availability

HCFC-22 CFC-12 Current Commercially available, widely used refrigerant.

CFC-502 Contains chlorine, will be phased out under

Montreal Protocol Copenhagen Amendments

HFC-134a CFC-12 Current Commercially available, rapidly expanding

production

Ternary CFC-12 (w/HCFC-124) Available in limited amounts

1993

HFC-152a CFC-12 Current Commercially produced and sold in fairly small

quantities, used primarily as a component in

CFC-500 (26%) and as a component in aerosol

propellant blends

HFC-123 CFC-11 1990-1991 Toxicity tests have shown sufficient toxicity to set

AEL at 10 ppm; commercially available

HFC-124 CFC-114 1993-95 Co-product of HCFC-123 production. Long term

toxicity testing started

HFC-125 CFC-502 Available in Near-term availability in blends to replace CFC-502

limited amounts

HCFC-141 b CFC-11 Current Commercial production began in July 1988.

Toxicity testing underway. Possible use as a foam

blowing agent

HCFC-142b CFC-12, 114 Current Used in R221R142b blends

NHs CFC-11 Current Commercially available. Widely used in industrial

CFC-12 refrigeration sector. Toxic with low flammability

CFC-502

Source: Arthur D. Little, 1993

The 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

variabl~ 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 altematives, this review of the technology of small

hermetic refrigeration compressors was undertaken as part of a larger study to evaluate

2-3

 CE-RM-93-801 COMMENT OR4
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 RIP energy

consumption reductions that can be obtained through the use of high efficiency small

compressors.

2-4

 CE-RM-93-801 COMMENT OR4
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 R/F units, in the United States

and worldwide:

.. Reciprocating

.. "Rotary" (Rolling Piston - stationary vane)

Both types are welded hermetic, i.e. the compressor pump and its motor are sealed

inside a welded shelL The refrigerant gas connections are welded to the shell and the

electrical connections to the motor are made through an insulated pass through bonded

to the metal shell. This hermetic arrangement prevents the loss of oil and refrigerant

which could occur through rotating seals or mechanical fittings.

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 he

slugged into the cannister on start up rather than into the cylinder. This could damage

the suction valve on the compression stroke. The liquid is vaporized harmlessly by the

3-1

 CE-RM-93-801 COMMENT OR4
Figure 3·1: Typical Small Hermetic Reciprocating Motor-Compressor

Suction

Discharge

/

crankShaft

Discharge

Nuffler

Cavity

.. -

M

\.l-IIIU~.11 HIT''+IIH'II" -'==" I

;: ....... 1-1111-1--,+-1-+-+-+-4·-.

.-.. nl

Piston &

Connecting

Rod

Discharge

Valve & Port

Suction Valve

& Port

Electrical

[ Conne"or

Mounting

Springs Motor

Rotor

OJ.l Pump

Inlet

'- Oil Sump

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

3-2

 CE-RM-93-801 COMMENT OR4
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 RJF 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 RJFs 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

 CE-RM-93-801 COMMENT OR4
Figure 3-2: Typical Small Hermetic Rotary Motor-Compressor

Rolling

"Piston

Crankshaft

Eccentric

Oil Inlet

To Shaft

Port

Motor Rotor

Motor Stator

&Windings

Illflll~l-suctionPort

Sliding Valve

and Spring

Suction Port

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

3-4

 CE-RM-93-801 COMMENT OR4
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 fractions 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

 CE-RM-93-801 COMMENT OR4
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 rpm; 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 tum reduces motor current, decreasing winding resistance losses

and improving motor efficiency by about 10%. This is called a capacitor run (CR)

3-6

 CE-RM-93-801 COMMENT OR4
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 CUlTent employed, approximately one half to one third of current

averages:

• Present capacity range - 500 to 1500 Btufhr

• 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 Btufhr range is available to U.S. RtF manufacturers (Table 3-2). As shown by the

data in Table 3-2 and graphically in Figure 3-3, an efficiency decrease parallels the

decrease in size, as indicated by the EER* value.

In Figure 3-3, several curves are shown to fit the individual EER data points. Somewhat

arbitrarily, below 600 Btu/hr is labeled "small compressor", above is labeled "large

compressor." CUlTent 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: BtulWatt-hr).

This expression is a variation of the "Coefficient of Performance" (COP) but well adapted to the units in

general use for expressing compressor data. The standard conditions for measuring the terms of this

value are -10°F saturated suction temperature, 130°F saturated discharge temperature, gO°F liquid and

suction vapor temperatures and 90·F ambient temperature.

3-7

 CE-RM-93-801 COMMENT OR4
Table 3·1: RefrigeratorlFreezer Manufacturers and Compressor Suppliers

Whirlpool

hite/Frigidaire

mana

Admiral/Maytag

Source: References 4, 24

Compressor Suppliers

General Electric

Danfoss

Tecumseh

Panasonic (Matsushita)

Embraco

Aspera

Americold

Panasonic

Sanyo

Tecumseh

Panasonic

Sanyo

Embraco

Tecumseh

3-8

 CE-RM-93-801 COMMENT OR4
able 3-2: Partla"I L"1st 0 f Current H"IgihEft"IClency . Aef .r1geratorIFreezer Compressors

Manufacturer Cap Dspl EER Motor Cooling Type

Sanyo (C030) 139 0.109 2.32 RSIR Static Recip

Sanyo (C-M40L12C) 180 0.13 2.6 RSIR Static Recip

Tecumseh (AZ1320D) 200 0.136 3.13 RSIR Static Recip

Embraco (EMI 20ER) 220 0.139 3.6 RSIR Static Recip

Panasonic (S070LKAA) 278 0.184 3.1 RSIR Static Recip

Embraco (EM 30SC) 315 0.184 3.89 RSCR Static Recip

Panasonic (S090LKAA) 349 0.217 3.5 RSIR Static Recip

Embraco (EM 40SC) 420 0.23 4.3 RSIR Static Recip

Americold (ST104) 440 0.217 4.7 PTCRICR Fan Recip

Embraco (EMI 40ER) 440 0.23 4.4 RSIR Static Recip

Panasonic (D112LRAA) 476 0.264 4.4 RSCR Static Recip

Panasonic (DA43L67) 516 0.264 4.75 RSCR Static Recip

Embraco (EM 55SC) 565 0.305 4.41 RSCR Static Recip

Embraco (EMI 55ER) 580 0.305 4.5 RSIR Static Recip

Americold (ST1 05) 580 0.277 5 PTCRICR Fan Recip

Americold (HG106-1 ) 621 0.277 5.35 PTCRICR Fan Recip

Panasonic (DA51 L88R) 635 0.311 5 RSCR Static Recip

Panasonic (FN40R80R) 651 0.248 4.46 RSCR Static Rotary

Americold (ST106) 660 0.312 5.06 PTCRICR Fan Recip

Americold (HG1 07-1) 708 0.312 5.4 PTCRICR Fan Recip

Tecumseh (AE1370W) 740 0.421 4.6 PTCRICR Static Recip

Panasonic (RA48L83R) 754 0.294 4.86 RSCR Fan Rotary

Panasonic (DA66L11 R) 806 0.402 5.15 RSCR Fan Recip

Americold (ST107) 810 0.366 5.06 PTCRICR Fan Recip

Americold (HG108-1 ) 848 0.361 5.5 PTCRICR Fan Recip

Panasonic (RA53L11 R) 850 0.323 4.8 RSCR Fan Rotary

Americold (HG109-1 ) 947 0.401 5.53 PTCRICR Fan Recip

Panasonic (DA73L13R) 973 0.469 5.25 RSCR Fan Recip

Panasonic (FN60R12R) 1000 0.372 4.61 RSCR Fan Rotary

Americold (HG110-1 ) 1029 0.473 5.51 PTCRICR Fan Recip

Panasonic (FN70R16R) 1151 0.435 4.79 RSCR Fan Rotary

Americold (HG111-1) 1193 0.5 5.55 PTCRICR Fan Recip

T

CAP

DSPL

RECIP

RSIR

PTCR/CR

RSCR

PSC

Capacity (Btu/hr)

Displacement (cubic inches) Reciprocating

Resistance Start Induction Run

Positive Temperature Coefficient Resistor

Start/Capacitor Run

Resistance Start Capacitor Run

Permanent Split Capacitor

Source: Engineering data supplied by the manufacturer of each model

3-9

 CE-RM-93-801 COMMENT OR4
Figure 3-3: Compressor Efficiencies versus Capacity

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

5.5

Current, State-of-the-Art

Large ReCiPs~ CFC·12 Refrigerant ______________________________________ u • .__ ...A. ~--n---A-- _

A fA

5

4.5

... __ u ••• __ •• ...... ...__n._~ n __ ..__ n •• n - _

III ilL IiIl • • • IiII _________ • ..... __ • __ ._. • ••.. u _

e. Current, State-of-the-Art

Large Rotaries

II:

UJ 4 UJ

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

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

•- - __ - - - - • - __ n n •• n _

3.5

3

• •

Current, State-of·the-Art

·Small" Compressors

Static Cooled •

Fan Cooled, Recip ...

Fan Cooled, Rotary ..

2.5

•

200 400 600 800 1000 1200

2'------......-------L-.-----.L------'-------L.-- -'--_--'

o

Cooling Capacity, Btulhr

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

 CE-RM-93-801 COMMENT OR4
Figure 3-4: Estimated Price of Average Efficiency Refrigeration Compressors

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

400 600 800 1,000 1,200

Nominal Capacity, Btu/hr

200

O'-------'-------'-------L----------J'-------'--- -'

o

30

10

20

3.4 Losses

As discussed above, the maximum efficiency level that is available in cUlTent,

production compressors, in terms of standard rating conditions EER, is 5.0 BtulWatt-hr

for "large" (nominal capacity above 700 Btufhr) and 3.5 for the small (200 Btufhr)

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

 CE-RM-93-801 COMMENT OR4
Figure 3-5: Ideal PV Processes at Standard Conditions

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

200 ..........................................................•..................................................................

.~ 150 0.

I

\

I

I

\

\ -----~._-,--- --_.--._-----..------.._------------..._--------------------------.-------_.-----.-_.._---------.._----.....--- I

\

\

\

\

\

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

"""'" " 50 :::::~

1.5 2 2.5 3

Specific Volume It Ilbm

Curve A Curve B

0.5

oL.-----'-----'------'-----'-----'----..........l

o

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-l2, the refrigerant that is the basis for all current

3-12

 CE-RM-93-801 COMMENT OR4
Condenser

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

Cooling

Pass~s ) ~~~ilillllil~ ........... I > I (

S",'on _ - I'---..:~~c~om~pr~e~ss~o~r~~~~~{~~_-_-_-_-_~jl--" (

Gas - Compo

in

Discharge

Gas

r

I I

Liquid to capillary

tube/interchanger

Motor

compressors. Table 3-3 summarizes the ideal EER for CFC-12, HFC-134a, HFC-152a,

and cyclopropane; at standard conditions the difference in ideal EER between these

t1uids is small.

Table 3-3: EER Limiting Cases, Based on Alternate Compression Processes

Ideal EER'

Refrigerant Reversible Adiabatic Isothermar 130°F

CFC-12 9.26 9.8

HFC-134a 9.32 9.8

HFC-152a 9.26 9.8

Cyclopropane 9.15 9.7

At standard rating conditions, 100% efficient motor and thermodynamically reversible compression

process.

.. Curve B in Figure 4-1

Table 3-4 presents, for several 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

 CE-RM-93-801 COMMENT OR4
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

High Efficiency Small, low Typical Rotary

Reciprocating Efficiency (EER =4.7)

(EER =5.0) Reciprocating

(EER =3.2)

Input Electric Power 100% 100% 100%

Motor Loss 17% 28% 22%

Mechanical Losses 8% 10% 10%

Suction Gas Heating 12% 16% 3%

Discharge Port Losses 2% 2% 4%

Low Side Pressure Losses 4% 4% 2%

Compression and Expansion 2% 4% 3%

Losses

Piston Blow By/Internal Leakage 1% 1% 5%

Power to Compressor - Gas 54% 35% 51%

Delivered··

3.4.1 Motors

Table 3-5 shows the efficiencies currently attained with compressor motors now in use

for the range 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 RJFs 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

 CE-RM-93-801 COMMENT OR4
Table 3-5: Motor Efficiencies

Compressor Motor Today's Motor Possible

Rating HP Type Efficiency Type Efficiency

200 Btu 1/16 RSIR 70-73% PSC 84%

400 Btu 1/8 RSIR 73-76% PSC 85%

600 Btu 1/6 RSIR&PSC 78-82% PSC 86%

800 Btu 1/4 PSC 80-84% PSC 86%

Legend

RSIR Resistance Start Induction Run

PSC Permanent Split Capacitor

These improvements could result in motor costs from 1.5 to 3 times present costs, an

increase over current motor costs 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 playa 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.

3.4.2.1 Mechanical Efficiency. Bearings comprise the major elements that determine

the mechanical efficiency of a pump. They consist ofjournal 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

 CE-RM-93-801 COMMENT OR4
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

 CE-RM-93-801 COMMENT OR4
l

t

iJ

9

~.'~

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.

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 have a larger

displacement than if its volumetric efficiency were 100%.

Volumetrjc 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). Fmther, 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 now 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

 CE-RM-93-801 COMMENT OR4
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 230"F

(-10"F Saturated Suction, 130"F Saturated Discharge, Isentropic Compression)

000

ODO

400

300

2IlO

'00

'"' '"'

40

3D

20

10

0.0

6.0

4.0

10

2.0

1.0

0.0

0.6

0.4

0.3

0.2

0.1

·20

I

I

I

I

I

I

I

I

I

I

I

I

I

I

I

I

I

I

I

-I

6h =29Btullb

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

Some 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

 CE-RM-93-801 COMMENT OR4
Figure 3-8: Compressor Heat Distribution

Suction--.

Gas

Reciprocating Compressor

~~=----"IO>II=- Discharge ,- Gas

a motor current

Rotary Compressor r--- Discharge

a motor current 11 Gas

Oshell cooling

L---L Suction

Gas

manifold in the cylinder head, immediately adjacent to the high temperature discharge

manifold. Finally the gas passes into the cylinder, whose walls have been 1:leated to a

temperature intermediate to the inlet and discharge gas temperatures. As a result of

these multiple heating steps, the suction gas 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

 CE-RM-93-801 COMMENT OR4
Figure 3-9: Low Side Compressor

Suction

Muffler

Baffle

Oil Outlet

Suction

Chamber

Discharge

Chamber

Compressor

Valve Head

Joint Spring

Suction Tube

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

Suction Gas L,....--t-L--'----

Figure 3-10: Effect of Suction Gas Temperature on Compressor Work for Constant Mechanical &

Cylinder Efficiency

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

oLJ.<:.- --l '--- .l- -'-- --'- --L.J

90 120 150 180 210 240

Cylinder Suction Gas Temp Degree F at Beginning of Compression

270

3-20

 CE-RM-93-801 COMMENT OR4
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 I 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

 CE-RM-93-801 COMMENT OR4
Figure 3-11: Low Side versus High Side Crankcase in a Single Cylinder Compressor

P2 P1

Discharge

Gas

-8- P2 P1

I Suction

Gas

Compressor Shell

P1

P1

Suction

Gas

(during suction stroke)

P2

-8-I

Compressor Shell

P2

Discharge

Gas

3-22

 CE-RM-93-801 COMMENT OR4
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

BtulWatt-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

RefrigeratorlFreezers over the Past Decade (EER of Isentropic Compressor with 100%

efficient motor is 9.3 BtulWatt-hr)

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

5.5 - ------- - --------..------------.••-.--..-----·"199XPrll-Protiactl'oIfSOA-----··--------------··-·---------

1990 Production SOA

5 - ..-------------...-.....-----........----..-----------------------..--.

4.5 - •.-----------.....--...-------..----------------------------------------------.-.•.--.•--.--------••-..••----------.------.---.

§.

a:

w

w 198D Production SOA

4 -.--------------••-.-----.....--•••-..--.-----.-.---.----------.-.....-...--- -----...--......----••--.--------

3.5 - ---...-- -------••-------....----••---.----------..-------.---------"-------.--.-••--------------- -----------------.-------....

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

600 800 1000

Cooling Capacity (Btulhr)

1200 1400

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 pert'ormance 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

 CE-RM-93-801 COMMENT OR4
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

1980 Production Current Current Future Goal

State of the Art Production Production (EER =6.5)

(EER =4.0) (EER =5.0) State of the Art

Samples

(EER =5.3)

Input Electric Power 100% 100% 100% 100%

Motor Loss 23% 17% 16% 8%

Mechanical Losses 9% 8% 7% 6%

Suction Gas Heating 15% 12% 11% 8%

Discharge Port Losses 2% 2% 2% 2%

Low Side Pressure Losses 4% 4% 4% 3%

Compression and Expansion 3% 2% 2% 2%

Losses (clearance gas expansion)

Piston Blow By/Internal Leakage 1% 1% 1% 1%

I Power Delivered to Gas 43% 54% 57% 70%

I Compression

Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet

conditions to discharge pressure.

4.1.1 Non-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

 CE-RM-93-801 COMMENT OR4
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

Typical Rotary Reference (38) With Best Motor

(EER = 4.7) (EER=5.1) (EER = 5.6)

Input Electric Power 100% 100% 100%

Motor Loss 22% 22% 16%

Mechanical Losses 10% 8% 9%

Suction Gas Heating 3% 3% 3%

Discharge Port Losses 4% 31/2% 31/2%

Low Side Pressure Losses 2% 2% 2%

Compression and Expansion 3% 21/2% 21/2%

Losses

Piston Blow By/Internal Leakage 5% 4% 4%

Power to Compressor - Gas 51% 55% 60%

Delivered*

Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet

conditions to discharge pressure.

The 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

 CE-RM-93-801 COMMENT OR4
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

18

16

14

l!! 12 .!11

(5

0

of

0 10 U

UJ

0

"iii 8

CQ)

E

u

E 6

Base EER =4.7

4

2

0

4.5 5 5.5 6

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

 CE-RM-93-801 COMMENT OR4
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 Btulhr 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 Btulhr 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 Btulhr, 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

Btulhr 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

 CE-RM-93-801 COMMENT OR4
Table 5-1: Estimated Small Reciprocating Compressor Limiting Efficiency and Input Power

Distribution, Based on Loss Scaling

large loss Scaling loss Scaling limiting

Reciprocating Factor** Factor @ 1/4 Efficiency @

at limiting Capacity*** 200 Btu/hr

Efficiency

(EER = 6.5)

Input Electric Power 100% N/A N/A 100%

Motor Loss 8% Table 3 - 5 +2% 10%

Mechanical Losses 6% Capacity -1/6 1.26 7.5%

Suction Gas Heating 8% Capacity -1/6 1.26 10%

Discharge Port Losses 2% -- -- 2%

Low Side Pressure Losses 3% -- -- 3%

Compression and Expansion 2% Capacity -1/3 1.6 3.2%

Losses

Piston Blow By/Internal Leakage 1% -- -- 1%

Power Delivered to Gas 70% N/A N/A 63%

Compression*

EER 5.7

***

Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet

conditions to discharge pressure.

Scaling factors are discussed in 5.4

Comparison of 200 Btu/hr to 800 Btu/hr nominal capacities

Table 5-2: Small (200 Btu/hr Nominal Capacity) Reciprocating Compressor Input Power

Distribution at Several Efficiency levels

low Efficiency, Current Intermediate Efficiency Maximum Efficiency

Production SOA (EER = 4.2) EER =5.7

(EER = 3.2)

Input Electric Power 100% 100% 100%

Motor Loss 28% 18% 10%

Mechanical Losses 10% 10% 8%

Suction Gas Heating 16% 16% 10%

Discharge Port Losses 2% 2% 2%

Low Side Pressure 4% 4% 3%

Losses

Compression and 4% 4% 3%

Expansion Losses

Piston Blow By/Internal 1% 1% 1%

Leakage

Power Delivered to Gas 35% 45% 63%

Compression*

Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet

conditions to discharge pressure.

5-2

 CE-RM-93-801 COMMENT OR4
5.2 Rotary Compressors

Table 5-3 is an estimate of the maximum efficiency level that is attainable in a small,

200 Btulhr 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 Btulhr, EER 5.6) discussed in Section 4.2. The resulting EER of

4.4, which is considerably less than the estimated limiting efficiency for small

reciprocating, is indicative 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

"Large" Rotary Scaling Loss Scaling Small Rotary

at Limiting Factor·· Factor @ 1/4 Limiting

Efficiency Capacity··· Efficiency @

(800 Btulhr, 200 Btulhr

EER =5.6)

Input Electric Power 100% N/A N/A 100%

Motor Loss 16% Table 3 - 5 +2% 18%

Mechanical Losses 9% Capacity -1/6 1.26 12%

Suction Gas Heating 3% Capacity ·1/3 1.6 5%

Discharge Port Losses 31/2% -- -- 31/2%

Low Side Pressure Losses 2% -- -- 2%

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

Losses

Piston Blow By/Internal Leakage 4% Capacity -2/3 2.5 10%

Power Delivered to Gas 60% N/A N/A 47%

Compression"

EER 4.4

Based on reversible adiabatic compression of actual delivered mass flow rate from shell inlet

conditions to discharge pressure.

Scaling factors are discussed in 5.4

Comparison of 200 Btu/hr to 800 Btu/hr nominal capacities

5.3 Estimated Cost of Improved Efficiency Small Compressors

As discussed in Section 3.3, manufacturers are generally reluctant to discuss their OEM

pricing policies in any detail, citing both domestic and foreign competition and

confidential relationships with 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

 CE-RM-93-801 COMMENT OR4
Figure 5-1: Estimated Incremental Cost of Improved Efficiency Small Compressors

3.5 4 4.5 5 5.5

OL----'''''-----J'-- --' --' --L _

3

20

I!!

.!!!

"0

Q

15 i

0

()

::ll

w

0

iii

C.. 10 E

e

u

.E

5

25

5.4 Approach to Scaling of Losses

To estimate the potential efficiency level 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

 CE-RM-93-801 COMMENT OR4
Table 5-4: Scaling of losses in Reciprocating and Rotary Compressor

Scaling Rule for loss vs. Capacity

Reciprocating Rotary

Motor Loss Small - Table 3-5 Small Table 3-5

Mechanical Losses Capacity -1/6 Capacity -1/6

Suction Gas Heating Capacity-1/6 Capacity -1/3

Discharge Port Losses No Scale Effect No Scale Effect

Low Side Pressure Losses No Scale Effect No Scale Effect

Clearance Gas Reexpansion Capacity -1/3 No Scale Effect

Piston Blow By/Internal Leakage No Scale Effect Capacity -2/3

Motor losses. The treatment in Section 3.4.1 is the basis for relating losses to

capacity, with Table 3-5 summarizing the relationship between capacity and motor

efficiency.

The maximum motor efficiency for 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 ofscale is probably small, and is neglected in this treatment. This is the

one area where the scaling factors are favorable to small capacities.

5.4.1 Reciprocating Compressors

Mechanical Losses. Mechanical losses include shaft bearing friction losses and

piston-cylinder wall sliding friction (hydrodynamic bearing oil film shear stress in both

cases). Other, 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 oc diameter

Piston sliding surface area oc 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 . . d4/d1 d . 1/3 to mput power IS . = = capacIty

5-5

 CE-RM-93-801 COMMENT OR4
.. Shaft bearing friction: for a given journal bearing lubricant viscosity, LID 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 = dol = capacity"1I3.

.. The net effect of the relatively small piston friction plus the bearing friction is

estimated to be: mechanical power losses/total input power = capacity"1I6.

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: capacityo

1l6.

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 dolor capacity to 0

113 •

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

 CE-RM-93-801 COMMENT OR4
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

capacitil/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 on 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-l =capacitil/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 to

the perimeter which is proportional to d. The ratio of the loss to the capacity is d/d3 = d-2=capacity"2/3.

5-7

 CE-RM-93-801 COMMENT OR4
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 Comment Annual En

kWh*

4.85** Readily available 685

5.0 Best production 669 16

compressor

5.5 1993 production 618 67

6.0 Approaching practical 570 115

limits

3.6/4.7 Currently available 602 83

4.6/5.1 With efficient motor 537 148

5.3/5.5 Approaching practical 498 187

limits

At DOE test conditions, taken to be representative of actual use

** Annual energy savings relative to large compressor/single evaporator baseline

" Cabinet volumes (cubic feet):

freezer 4.6

fresh food 13.4

total 18.0

adjusted volume 20.9

" Federal energy efficiency standard:

1990 962 kWh/yr

1993 690 kWh/yr

" 4.85 EER, 893 Btulhr compressor (readily available, moderate cost, reasonable

efficiency, provides typical pulldown capacity).

" Cabinet insulation is polyurethane foam with R=8°F/in per Btulhr-ft2

average freezer wall thickness: 23/8 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

 CE-RM-93-801 COMMENT OR4
• 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 comprerssor 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 8¢ 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

 CE-RM-93-801 COMMENT OR4
Table 6-2: Present Value of Energy Savings Obtained with Increased Efficiency Compressors

Present Value for Discount Cost of

Rate Improvements

Compressor Savings 2% 5% 10% (in U.S. Dollars

EER kWh/yr** at Retail)

Large Compressor 4.85* -- -- -- -- 0

Single Evaporator 5.0 16 17 14 10 5

5.5 67 70 58 43 10

6.0 115 120 99 73 20

Small Compressors 3.6/4.7 83 87 71 3 20

Dual Loop 4.6/5.1 148 155 127 94 40

5.3/5.5 187 196 161 169 80

Baseline

** Relative to baseline

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7.0 Areas for 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 efficiency

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 RIP'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 savings

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 small compressor designs to significantly reduce suction gas

superheating. Other loss areas should also be systematically examined for 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 BtulWatt-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 RJFs 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

RlF 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.

2. Federal Register, November 17, 1989b, pp. 47916-47938.

3. Dieckmann, J. and McMahon, E., State of the Art Survey of 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, 1.

Hammond, T. Kandpal and HU Richardson at Tecumseh.

5. Robert Chilton, Danfoss, Inc., Mahwah, NJ, telcon 11/16/89.

6. Robert Hawley, VP Marketing Americold, White Consolidated Industries, Cullman,

AL, telcon 11/22/89.

7. 1. Vickerman, 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 aI, "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 CommerciallIndustrial Refrigeration Systems," Technical

Memorandum to U.S. DOE by Arthur D. Little, Inc., August 1980.

14. Nelson, R.T. and MacCarthy, P.W., Research and Development of Energy-Efficient

Appliance Motor-Compressors, Final Report, Vol. I Executive Summary and Vol.

III 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 RlF compressor field.

17. Laercio Hardt, Embraco, 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

Americo1d does not have compressors with EERs in the 5 range for the lower

capacities.

22. 1. 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 RlF 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 RlF sizes, not optimistic.

26. Mark Okazaki, Matsushita (Panasonic), US rep, telcon 12/12/89, obtain information

on Panasonic compressors, current and future plans.

27. 1. Hackstedde, Tecumseh Products Co., telcon 11/30/89. Tecumseh makes most

motors, buys some from GE.

28. Donald Allen, 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 RlF 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 aI, "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 aI, "Electronically Commutated Motors for the Appliance Industry,"

1984 IEEE publication.

37. "Recent Technological Trends and Lineup of Hitachi Compressors," K. Kikuchi, et

aI, 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."

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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