Investigation of the Effects of Heat Transfer from a Scroll Compressor through the Use of Heat Pipes

INVESTIGATION OF THE EFFECTS OF HEAT TRANSFER FROM A SCROLL
COMPRESSOR THROUGH THE USE OF HEAT PIPES
















By

KHIEM BAO DINH


A THESIS PRESENTED TO THE GRADUATE SCHOOL
OF THE UNIVERSITY OF FLORIDA IN PARTIAL FULFILLMENT
OF THE REQUIREMENTS FOR THE DEGREE OF
MASTER OF SCIENCE

UNIVERSITY OF FLORIDA


2005

































Copyright 2005

by

Khiem Bao Dinh

































This document is dedicated to my family and friends.















ACKNOWLEDGMENTS

I would like to take this opportunity to recognize the individuals that have

allowed me to perform this work. Foremost, I would like to thank Dr. Vernon Roan for

giving me the opportunity to attend the University of Florida and work in the fuel cell

lab. I have learned a great deal from Dr. Roan and his experience and I am grateful for

having had the opportunity to work for a person of the highest caliber.

I would also like to thank Robert Shaffer of Air Squared, Inc., and Khanh Dinh of

Heat Pipe Technology, Inc. Without the assistance of Air Squared in loaning the scroll

compressor, this research would never have been possible. In addition, Heat Pipe

Technology provided invaluable assistance in constructing the heat pipe.

My colleagues Daniel Betts, Timothy Simmons, and Alex Burrows proved

instrumental in helping me complete my research. Daniel and Timothy provided

invaluable advice and knowledge during the entire course of this research and Alex

provided much welcomed and needed assistance in the setup of the experiment.

Next, I would like to recognize the support from the University of Florida and the

Mechanical and Aerospace Engineering Department that has provided the resources to

further my education. I would especially like to thank Becky Hoover and Pam Simon for

making sure I registered for classes, filled out forms, turned in or signed any required

paperwork on time, and for just having someone to visit and talk with. In addition, I

would like to thank my thesis committee, Dr. William Lear and Dr. Skip Ingley, for their

support.









Finally, I would like to thank my family for their life-long support in my

endeavors. My parents Khanh and Hong raised me to perform to a higher standard in all

aspects of life and provided the support necessary for an unparalleled upbringing. My

older sisters Mai and Tina kept an eye on me while growing up, provided exceptional

examples of how to live life, and have always been there to help me in times of need.
















TABLE OF CONTENTS

page

A C K N O W L E D G M E N T S ................................................................................................. iv

L IS T O F T A B L E S ....................................................................... .............. ................... v iii

LIST OF FIGURES ......... ......................... ...... ........ ............ ix

ABSTRACT ........ .............. ............. .. ...... .......... .......... xii

CHAPTER

1 INTRODUCTION ............... ................. ........... ................. ... ..... 1

2 REVIEW OF LITERATURE ......................................................... .............. 5

Scroll C om pressor ....................................................... 5
H eat Pipe Background ...................................... ........................... .6

3 EXPERIMENTAL COMPONENTS AND SETUP .................................................8

Stock C om p ressor S etu p .................................................................... .......... .. .. ...8
H eat Pipe Integration .................. .................................... ................. 15

4 EXPERIMENTAL TEST PROCEDURE AND DATA ANALYSIS........................21

Sy stem C haracterization T ests ...................................................................................2 1
Experim ental Test Procedure......................................................... .............. 23
U n certainty ............................................................................2 7
Experim ental D ata and A analysis ........................................ .......................... 28
Motor Speeds...................................................... ... ........28
Compressor Component Temperatures .................................... ............... 28
Air Discharge Temperature and ATexit-in .................................. ...............30
Mass Flow Rate and Pressure Ratio ...................................... ...............31
P ow er C onsum ption......... .......................................................... ....... .... .... ....... 34
Effective System Efficiency ........................................ .......................... 38
A vailability-B ased Efficiency ........................................ ......... ............... 41
Approximated Compressor Efficiency ..................................... ............... 44









5 AIR COM PRESSION M ODEL ........................................ ........................... 47

M odeling P rocess........... ...... ............................................ .............. ......... ........ 4 7
M odel R results and A analysis ............................................... ............................ 51
T em perature Profi les ................................................. ............... ............... 52
Experimental Data and Model Results Comparison .......................................55
Power Distribution and Heat Removal Rate ................................................. 55

6 CON CLU SION S .................................. .. .......... .. .............62

APPENDIX

A EXPERIMENTAL COMPONENTS .............................. 65

B EXPERIMENTIAL RESULTS AND UNCERTAINTY.........................................68

C MODEL SPREADSHEET SAMPLE..................................... ........................ 75

D M O D E L G R A P H S .......................................................................... .....................77

E N O M E N C L A T U R E ......................................................................... ....................84

L IST O F R E FE R E N C E S ....................................................................... ... ...................87

B IO G R A PH IC A L SK E TCH ..................................................................... ..................88
















LIST OF TABLES

Table p

1 D ata averaging process sam ple ........................................... ......................... 27

2 A actual test speeds .................. ........ ................... ........... 28

3 Compressor component temperatures ........................................... ............... 29

4 Air intake and discharge temperatures and ATexit-in ............... ..... ...............30

5 M ass flow rates and pressure ratios...................................... ........................ 32

6 Calculated pow er values.......................................................... ............... 34

7 Calculated MC from trendlines for tested range ofPq-0oo .................................. 38

8 Effective system efficiencies, sy .................................... ................................39

9 Calculated ly from trendlines for tested range of motor speeds ............................40

10 Calculated rsy from trendlines for tested range of mass flow rates........................41

11 Calculated availabilty-based efficiency ....................................... ............... 42

12 Calculated rlA from trendlines for tested range of motor speeds............................43

13 Calculated rlA from trendlines for tested range of mass flow rates .........................44

14 Appoximated compressor shaft power and lsp .....................................................45

15 Approximated model first stage air temperatures and densities........................ 54

16 A T exit-wall values .......................................................................54

17 Comparison of experimental data and model results ............................................56

18 M odel predicted power distribution ..................................................................... 58

19 Model calculated heat transfer coefficient h .......................................................61
















LIST OF FIGURES

Figure page

1 Reduced compressor work due to cooling between stages .......................................3

2 Com pression process of a scroll com pressor ........................................ ..................6

3 Heat pipe schematic ................... .. ........... ... ............ .. ............. .. 7

4 Com pressor system ...................... .... ............ ............................ .8

5 Com pressor assem bly ................................................ ... ...... .. ........ .. ..

6 Compressor housing and counterweighted motor shaft .................. ...............

7 F ix ed scro ll ...................... .. .. ......... .. .. ......... .................................... 10

8 Orbiting scroll and scroll/m otor shaft bearing ............................... ............... .10

9 Air filter and intake temperature thermocouple .................................................11

10 14 NPT-to-hose barb fitting with thermocouple ................................................. 12

11 Thermocouple holes in fixed scroll wall ............................................................ 12

12 Fixed scroll inlet ports, exit port, fins and thermocouples................................ 13

13 Thermocouple on fixed scroll fin.......................... .............. ............... 13

14 Therm ocouple on m otor ........................ ... ..................................... ............... 14

15 Therm ocouple on m otor controller ............................................... ............... 14

16 A irflow path diagram ....................................................................... .................. 15

17 Grooves machined into fixed scroll between bearings and air inlet and exit ports..16

18 Depth of groove machined into fixed scroll...................................................17

19 H eat pipe assem bly ........... ................................................................ .. .... ... ... ... 17

20 Therm couple placed in heat pipe fins ........................................ .....................18









21 Heat pipe clamped to fixed scroll................ .......................... ... ............ 18

22 Interface of heat pipe and fixed scroll .............. ........................... .............. 19

23 Fans mounted to the heat pipe assembly ............... ....... ..................... 19

24 View of metal plates used to duct airflow .................. ................................. 20

25 Initial test tem perature graph...................... ................................. ............... 21

26 ATexit-in vs. M otor Speed ............. ............................ ...................... ............... 31

27 M ass flow rate vs. m otor speed ..................................................... ............... 32

28 Pressure ratio vs. m otor speed ...................................... ............................. 33

29 Pressure ratio vs. m ass flow rate ........................................ ......................... 33

30 M C vs. m otor speed ................................................ ...............35

31 P n 1oo0%vs. m otor speed ............................ .............. ... ...... .. .... ...........35

32 M C vs. m ass fl ow rate............................. ................ ...................... ............... 36

33 P n oo0% v s. m ass flow rate.............................................................. .....................36

3 4 M C v s P 100 ........................ ................................................................................. 3 7

35 r sys v s. m otor speed .......................... .. ........................ .. ........ ................39

36 rsys vs. mass flow rate....................... ................. ....... ......40

37 lA V S. m otor speed ............ ............................................................... .......... ....... 42

3 8 lA V S. m ass flow rate .................................. ...................................... .... 4 3

39 rsp vs. m otor speed ........................... ........ .. ...... ...............46

40 sp v s. m ass flow rate............ .................................. ............................... 46

41 Model predicted temperature profile for stock configuration at 900 rpm ..............53

42 Model predicted temperature profile for forced convection configuration at
10 8 0 rp m ............................................................................ 5 3

43 Model predicted Qdotremoved, total vs. motor speeds ................................................57

44 Model predicted Qdotremoved, total vs. mass flow rates.................................... 58

45 Model predicted Motoriosses vs. motor speed.........................................................59


x









46 M odel predicted M E1 losses vs. m otor speed ..................................... .................59

47 Model predicted Motoriosses vs. mass flow rate ................................ ...............60

48 M odel predicted M Elosses vs. m ass flow rate ................................. ................60















Abstract of Thesis Presented to the Graduate School
of the University of Florida in Partial Fulfillment of the
Requirements for the Degree of Master of Science

INVESTIGATION OF THE EFFECTS OF HEAT TRANSFER FROM A SCROLL
COMPRESSOR THROUGH THE USE OF HEAT PIPES


By

Khiem Bao Dinh

August 2005

Chair: Vernon P. Roan
Major Department: Mechanical and Aerospace Engineering

A scroll compressor is an efficient compressor design commonly used in air

conditioning and refrigeration applications. A relatively new application for compressors

is with fuel cell system pressurization in which they are used to increase the fuel cell

stack power output. The scroll compressor is one type of compressor being investigated

and integrated for use with these systems.

First developed by NASA, heat pipes are passive heat transfer devices with high

effective thermal conductivities and are now used in a wide range of common

applications that require the transfer of heat from one location to another. For example,

heat pipes are used in laptop computers to transfer heat from the microprocessor chip to a

fin assembly that is located more conveniently for packaging.

The purpose of this investigation was to provide information regarding the effect of

heat transfer from a scroll compressor by means of a heat pipe. A scroll compressor was

experimentally tested in three configurations: stock, integrated with a heat pipe rejecting









heat by free convection to the ambient atmosphere, and integrated with a heat pipe

rejecting heat by forced convection to the ambient atmosphere. Each configuration was

tested over a range of motor speeds. Furthermore, a simple computer model was

developed and used to further analyze the experimental data.

The results show that heat transfer from the scroll compressor through the use of

heat pipes has positive effects on increasing mass flow rates, reducing power

consumption and increasing efficiencies the compressor achieves. The information

presented in this thesis should be coupled with a cost and system integration analysis to

determine whether the use of heat pipes with scroll compressors would be beneficial.














CHAPTER 1
INTRODUCTION

Compressors are important to meeting the standard of living enjoyed by modern

society. The widespread use of vapor-compression refrigeration and air conditioning

around the world has made the compressor an indispensable device. The number of air

conditioning and refrigeration currently used numbers in the hundreds of millions, and

with the increasing development of third world countries even greater numbers of

compressors are being used.

A growing and potentially vast market for compressors lies with the increase of

fuel cells, a market projected to expand rapidly in the next few decades. Fuel cells are

viewed by many as the future of power generation as an alternative to fossil fuel

combustion and nuclear power generation. Two essential characteristics of fuel cells are

their high efficiency and environmental cleanliness, with their only emission being water.

The integration of compressors into fuel cell systems is being scrutinized since the

pressurization fuel cells provide a number of benefits.

The pressurization of a fuel cell through the use of a compressor causes the fuel cell

stack to have a higher efficiency and greater power density (desirable especially in

transportation applications). Furthermore, when a compressor is used in conjunction with

a Proton Exchange Membrane (PEM) fuel cell, the pressurization of the fuel cell aids in

the water management that is vital to the operation of a PEM fuel cell.3 The major

drawback of compressor use with fuel cells is the large power requirement for

compressor operation; the power drawn by the compressor may negatively impact the









overall fuel cell system more than the compressor aids the system. Therefore,

maximizing the efficiency of a compressor is critical in a fuel cell system.

In the simplest of terms, a compressor is a device used to increase the pressure of a

gas. For the compression of air, there are two general types of compressors: positive-

displacement and dynamic. Positive displacement compressors such as reciprocating and

rotary compressors increase the pressure of the air by decreasing its volume. Dynamic

air compressors use high velocity impellers to transfer angular momentum from the

impeller to the air thereby increasing the pressure of the air. Dynamic compressors

include axial flow and centrifugal compressors. In general, positive displacement

compressors are used for applications involving lower capacities (flow rates) and higher

pressure ratios whereas dynamic compressors are used for applications involving higher

capacities and lower pressure ratios.1

One desirable compression process would be an isothermal process where heat is

constantly removed during the compression process maintaining the gas at a constant

temperature. Isothermal compression reduces the work required to compress the gas

compared to compression processes where the gas temperature increases during

compression. This can be seen by the reduction of area under the pressure-volume curve

of isothermal compression versus polytropic compression in Figure 1. Very few real

compression processes are able to achieve isothermal compression, but multi-staged

compression processes have been devised where heat is removed between compression

stages to reduce compressor work.

A two-stage compression process with an intercooling intermediate step is also

shown in Figure 1. Intercooling is the removal of heat from a gas after being










compressed. In a staged compression system, the gas passes through one compressor

reaching a higher pressure, flows through a heat exchanger (intercooler) that cools the gas

to a lower temperature before passing through another compressor to reach a final

elevated pressure. The removal of heat in an intermediate step reduces the total

compressor work required versus compressing the fluid in one single step to the same

-5
pressure without intercooling.5

P1 T' P2
Work saved
p2 ..... / ..

SPolytropic P1



Intercooling 1
Isothermal Intercooling

1 1


U Z

Figure 1. Reduced compressor work due to cooling between stages

A compressor design commonly used in modern air conditioning units is the scroll

compressor. This compressor has a large, stationary surface area in contact with the

compressed gas during compression making it well suited to intercooling. The unique

geometry and operation of the scroll compressor allows for heat transfer during the

continuous compression process and internal to the compressor. Therefore, the heat

transfer during compression may allow for isothermal compression.

The intent of this study is to provide information on the effects of heat transfer

from a scroll compressor by means of an integrated heat pipe. It was hypothesized that

the main effect would be reduced power consumption for a given mass flow rate and

pressure ratio. A scroll compressor was tested over a variety of motor speeds in three






4


different configurations with each configuration having a different rate of heat transfer

from the compressor system. A computer model was also developed to better analyze the

experimental data in determining the performance of the compressor system.














CHAPTER 2
REVIEW OF LITERATURE

Scroll Compressor

In 1905, Leon Creux invented the scroll compressor, which is essentially a positive

displacement type compressor. The basis of the scroll compressor is two identical

involute spirals, or scrolls, that are nested together. With the two scrolls mated together,

they create a series of crescent shaped air pockets between them.4

The scroll compressor operates by keeping one scroll stationary (fixed scroll) while

the other scroll (orbiting scroll) orbits around the fixed scroll. The compression process

is shown in Figure 2. The first step is the intake of the air through the air inlets. The

second step is the sealing off of the air forming an air pocket. The orbiting scroll motion

causes the gas pockets to move towards the center of the scrolls while being reduced in

volume, hence the compression. Once the gas pocket reaches the center of the scrolls, the

tip of the orbiting scroll uncovers the discharge port located in the center of the fixed

scroll thereby beginning the discharge process. Due to the geometry of the scrolls, a

discharge valve is unnecessary as compared to a reciprocating type positive displacement

compressor. The discharge of air is nearly continuous as multiple pockets of gas are

compressed simultaneously (Figure 2, step 5).7











Aii inlets Orbiting
scroll



SAir pocket
1 2
Fixed
scroll



Ail dtclirge


3 4 5
Figure 2. Compression process of a scroll compressor

Heat Pipe Background

Heat pipes are passive heat transfer devices with high effective thermal

conductivities that are based on a closed two-phase cycle and use the latent heat of

vaporization to transfer heat. The two-phase cycle allows heat pipes to have a heat

transfer capacity greater than the best solid conductors by several orders of magnitude

and leads the heat pipe to being a nearly isothermal device. In comparing aluminum and

copper rods with a heat pipe (all of 0.5m length and 1.27cm diameter) transmitting

twenty watts of power, the aluminum has a AT of 460 C, the copper a AT of 206 C, and

the heat pipe a AT of 6 C.

Heat pipes typically consist of a sealed container with an internal wicking material

and working fluid (Figure 3) and can be broken down into three major sections:

evaporator, condenser, and an adiabatic/isothermal section in between. Heat addition

occurs at the evaporator where the working fluid in liquid phase is heated until it










vaporizes. The vapor then flows to the condenser and changes phase back to a liquid

releasing the latent heat of vaporization. Capillary forces in the wicking structure pump

the liquid back to the evaporator section.8


HEAT PIPE WALL
CAPILLARY WICK

LIQUID FLOW












HEAT IN HEAT OUT
EVAPORATOR CONDENSER


Figure 3. Heat pipe schematic

Changes in the heat flux into and out of the heat pipe results in a varied rate of

vaporization and condensation of the working fluid without significantly increasing the

operating temperature. Therefore, a heat pipe will operate at a relatively constant source

temperature over a wide range of power input by adjusting the rate at which the fluid

cycles through the heat pipe.

Next, the evaporator and condenser sections may be of different sizes and locations,

only needing common liquid and vapor streams. Therefore, high heat fluxes generated

over a small area (small evaporator section) can be dissipated over a larger area (large

condenser section) at a smaller heat flux. An additional advantage is the thermal

response time of a heat pipe; it's faster than solid conductors and is also not a significant

function of length.6














CHAPTER 3
EXPERIMENTAL COMPONENTS AND SETUP

Stock Compressor Setup

The main component of the experimental test setup was the scroll compressor. A

prototype scroll compressor system sourced from Air Squared, Inc. consisted of: a three-

phase brushless DC motor, a compressor housing, and a compressor assembly (Figure

4)1.


compressor
brushless assembly
DC motor










housing


Figure 4. Compressor system

The compressor housing attached to the motor acting as the mating adaptor for the

compressor assembly to the motor. The compressor assembly was composed of the fixed

and orbiting scrolls (Figure 5). A counterweighted shaft connected to the motor spins the

orbiting scroll (Figure 6).


1 Please see Appendix A for detailed compressor specifications
























Figure 5. Compressor assembly

A high efficiency brushless DC motor was utilized in Air Squared's system. The

compressor housing was vented to allow for airflow over the finned orbiting scroll for

heat rejection.


















Figure 6. Compressor housing and counterweighted motor shaft

The orbiting scroll was mated to the fixed scroll by means of three shafts that are

located by ball bearings in each scroll for reduced friction. This scroll compressor

required no lubrication and utilized tip seals between the mating surfaces (scroll wall and

endplates) of the fixed and orbiting scrolls (Figure 7).





























Figure 7. Fixed scroll

A ball bearing was used to mate the motor shaft to the orbiting scroll for reduced

friction (Figure 8) and two air inlet ports were used instead of one to reduce pumping

losses. These features all contributed to minimizing losses.


bearings eaIng


Figure 8. Orbiting scroll and scroll/motor shaft bearing









An adjustable, 30-volt and 5-amp power supply was used to power the motor

controller that in turn was used to power the motor.

The measurements of interest were: volumetric airflow rate, ambient air pressure,

compressor discharge air pressure, voltage and current input to the motor controller,

motor speed, and various temperatures. Airflow rate was measured using a digital

volumetric flow meter. Air pressure generated by the compressor was measured with a

liquid filled mechanical gauge and an indoor weather station measured ambient air

pressure. Digital multi-meters measured voltage, current, and motor speed.

Temperatures of interest were: air compressor intake and discharge, fixed scroll wall and

fin, motor, and motor controller. Temperatures were measured and data-logged using

Omega thermocouples and lotech data acquisition hardware and software.

Air compressor intake and discharge temperatures were measured using

thermocouples placed in fittings attached to the air filter and compressor exit port

(Figures 9 and 10).


Filter


Figure 9. Air filter and intake temperature thermocouple
























Figure 10. 14 NPT-to-hose barb fitting with thermocouple

The fixed scroll was modified to allow for measurement of the scroll wall

temperatures. Three holes were drilled into the scroll wall at locations reflecting the

beginning, middle, and end (thermocouples Scroll 1, 2, and 3 respectively) of the airflow

path through the compressor (Figures 11 and 12). The thermocouples were inserted

approximately 20mm into the hole from the endplate side, which positioned the

thermocouples roughly in the middle of the scroll wall by height, and sealed with RTV

grey gasket maker.


thermocouple holes


Figure 11. Thermocouple holes in fixed scroll wall


















Figure 12. Fixed scroll inlet ports, exit port, fins and thermocouples
Thermocouples were attached to a fin on the fixed scroll (Figure 13), the motor casing
(Figure 14), and motor controller (Figure 15) with adhesive backed thermocouples from
Omega.


I


Figure 13. Thermocouple on fixed scroll fin






















Figure 14. Thermocouple on motor


Figure 15. Thermocouple on motor controller

The airflow path began through the air filter. A thermocouple was placed in a

fitting attached to the filter to measure air intake temperatures (Figures 9 and 10). Next,

the air flowed through a 10 cm length of 9.5 mm I.D clear vinyl tubing to the volumetric

flow meter before flowing through 50 cm of tubing to a T-fitting. From the T-fitting, the

airflow split to two 25 cm sections of tubing that attach to the two intake ports of the

compressor. At the compressor exit port, a 14 NPT-to-hose barb fitting with a

thermocouple was screwed into the port allowing discharge air temperature measurement.

30 cm of 50 psi (345 kPa) rated fuel line was used on the compressor exit side, acting as

the discharge reservoir, because the vinyl tubing used on the intake side expanded under

pressure and elevated temperatures. A ball valve attached to the end of the fuel line acted

as a flow restriction and controlled the exit area of the flow. A T-fitting was placed 5 cm










before the ball valve as a pressure tap for the pressure gauge to measure discharge

pressure. Figure 16 is a path diagram for the airflow.


10cn length tube
-- Air filter Flow meter
50cm length
1 .'.- l I r tL bell s tubees t
Compressor


30cm length
tube
Pressure gauge

Ball valve



Figure 16. Airflow path diagram

The power used by the compressor system was determined by measuring the

voltage and current going to the motor controller. The motor and motor controller were

together considered to be an integrated system. Motor speed was determined by

measuring the frequency from a motor Hall effect sensor referenced to the ground of the

motor controller.

Heat Pipe Integration

The compressor required modification to accommodate the addition of a heat pipe

to the fixed scroll whose geometry was limiting in the configurations available. The

obstacles to fitting the heat pipe were the three ball bearings, air inlet ports, and the air

exit port. The heat pipe would have to go around these objects, as they could not be

modified. Therefore, this left two narrow paths between the bearings and air inlet and

exit ports where the heat pipes could be placed. Unfortunately, this resulted in the

removal of the third fixed scroll thermocouple (#3) that was located nearest the discharge

port.









The heat pipe diameter was the next variable to be determined. The maximum

diameter heat pipe that could be fitted was 12.7 mm diameter. An alternative option was

the use of two 6.35 mm diameter heat pipes in place of a single 12.7 mm diameter pipe to

increase the contact surface area between the heat pipes and fixed scroll. It was

determined that the heat transfer properties of the 6.35 mm heat pipes were relatively

limited and therefore, the 12.7 mm heat pipe was chosen. The larger diameter pipe also

reduced the machining necessary, thus easing the modification.

The first step in machining the fixed scroll was using a 12.7 mm diameter end mill

creating the paths for the heat pipe through the fins. Next, a circular groove of 6.35 mm

radius was machined using a 12.7 mm diameter ball-end end mill, resulting in the mating

interface between the heat pipes and fixed scroll surface (Figure 17).




















Figure 17. Grooves machined into fixed scroll between bearings and air inlet and exit
ports

The fixed scroll end plate was 5 mm thick limiting the possible depth of the groove

resulting in a machined groove depth of 1.78 mm (Figure 18). This depth created a







mating interface area totaling approximately 19.5 cm2 (2 grooves x 9.74 mm arc length x
100 mm groove length) while not affecting the structural rigidity of the fixed scroll.












Figure 18. Depth of groove machined into fixed scroll
The heat pipe was fabricated with the assistance of Heat Pipe Technology, Inc. and
had the form of a U-shape with a length of 61 cm. A 30.5 cm section of the condenser
side of the heat pipe was finned using aluminum fins of 25 mm width, 63.5 mm length,
and 4.33 fins per centimeter. Three thermocouples were attached to the heat pipe: one on
each side of the U (labeled Pipe 1,2) and one in the fin assembly (Pipe 3). Figures 19 and
20 show the heat pipe, fins, and thermocouples.


thermocouples

V<


0


Figure 19. Heat pipe assembly























Figure 20. Thermocouple placed in heat pipe fins

The heat pipe fit tightly into the machined grooves but a small clamp fitting was

fabricated to secure the heat pipe to the endplate (Figure 21). Thermal conducting paste

was used at the interface between the heat pipe and the endplate to ensure good heat

conduction (Figure 22).


Figure 21. Heat pipe clamped to fixed scroll
























Figure 22. Interface of heat pipe and fixed scroll

The forced convection tests required the addition of fans to the heat pipe assembly

to reject heat to the atmosphere. Four small fans generally used for cooling CPU chips in

computers were attached to the heat pipe assembly and powered by an additional power

supply. To ensure airflow through the fins, two metal plates were attached to the sides of

the fins to duct the airflow. Figures 23 and 24 show the fans and metal plates attached to

the fins of the heat pipe.


metal plates

- S A -


Figure 23. Fans mounted to the heat pipe assembly








V"


metal plates
Figure 24. View of metal plates used to duct airflow
















CHAPTER 4
EXPERIMENTAL TEST PROCEDURE AND DATA ANALYSIS

System Characterization Tests

A series of initial system characterization tests were performed to determine the

range of the measured parameters of power input, volumetric flow rates, pressure ratios

and thermocouple temperatures. The test plan required steady state conditions; therefore,

these initial tests were performed to determine the time required for the system to reach

steady state with respect to scroll and discharge air temperatures. Figure 25 shows an

initial test run and the time required for the temperatures to reach a steady state (up to 2

hours), defined as the mean temperature of each thermocouple reading varying by less

than 1 C over a ten-minute time interval.



Thermocouple Measurements

100
90
80
70
60
50
40
E 30- -
20
10
0
0 1000 2000 3000 4000 5000 6000 7000 8000
time (sec)

Intake scroll 1 scroll 2 Xscroll 3 exit +fin -motor controller


Figure 25. Initial test temperature graph









The initial tests exposed other characteristics of importance. Scroll 1, Scroll 2, and

Scroll 3 represent the three thermocouples placed into the fixed scroll wall (Figure 12).

Figure 25 shows Scroll 1 maintained the lowest temperature, Scroll 2 an intermediate

temperature, and Scroll 3 the highest temperature. It was hypothesized these

temperatures reflected the increasing air temperature as it was compressed along the

airflow path. Also, the discharge air temperature was directly related to the fixed scroll

wall temperature shown by the discharge air temperature increasing proportionally to the

fixed scroll wall temperatures.

Additional temperatures of importance were the fixed scroll fin, motor, and motor

controller. The fixed scroll fin and motor temperatures were also directly related to the

fixed scroll wall temperatures. The motor controller temperature was found to increase

similarly to the motor during an initial test and therefore, a fan was placed over the motor

controller to maintain a cool temperature.

Final observations from the initial tests relate to the volumetric airflow rate, motor

controller power input, discharge air pressure, and motor speed. The first three variables

were somewhat unsteady, given a steady motor speed. The motor speed fluctuated over a

range of 1.5 rpm at motor speeds of 960 rpm and 1020 rpm, less than 0.14% of full scale

(1080 rpm). Motor speed fluctuated over a range of 0.3 rpm at 900 rpm and 0.6 rpm

1080 rpm. The measured discharge air pressure oscillated apparently due to the nature of

the scroll compressor. Though the air pressure gauge was liquid filled, the reading

oscillated approximately 0.25 psi (1.7 kPa) over the range of 15-20 psig (103-138 kPa).

The current input to the motor controller varied up to 0.15 amps (over a range of 3-5

amps) given a constant voltage input at constant motor speed. Increased current at









constant motor speed resulted in increased air volumetric flow rate and discharge

pressure. Based on these system characteristics, motor speed and airflow exit orifice area

could be assumed to be constant during the test runs and the other measured variables

would be continuously monitored.

Experimental Test Procedure

A fixed ball valve position at the airflow exit, and four motor speeds (900 rpm, 960

rpm, 1020 rpm, and 1080 rpm) were used for all test runs. For each test run, the motor

was started and run at 900 rpm until the thermocouple temperatures reached steady state.

Data was then recorded once every minute until ten data points were collected. Next, the

motor speed was increased to 960 rpm and temperatures allowed to reach steady state

before recording another set of data points. This procedure was repeated for 1020 rpm

and 1080 rpm. The values recorded were: temperature readings (air intake and discharge,

Scroll 1-2, Pipe 1-3, fin, motor, motor controller), voltage and current to the motor

controller, volumetric air flow rate, ambient air pressure, and discharge air gauge

pressure from the compressor. These values were then used to calculate a number of

parameters. It is important to note that the parameters calculated using the measured

discharge air temperature do not represent truly accurate values because the measured

temperatures were dependent on thermocouple location; however, the parameters may

confidently be compared relative to each other as they all used data from the same

thermocouple and location.

Inlet air density from the equation of state for an ideal gas:


S *P=mb Eq. 1
R*Tn

M = molecular weight of air









Pamb = ambient air pressure

R = universal gas constant

T = compressor intake temperature

Mass flow rate:


Ih = V*p Eq. 2

V = volumetric flow rate

Pressure ratio:


PR= dp Eq. 3
Pamb
dp = discharge pressure = Pdischarge,gage + Pamb,abs


Net rate of heat addition:


Qne, = A M Cp *(T n) Eq. 4

Tex, = compressor discharge air temperature

Ah = change of enthalpy

C = average specific heat

Qet represents the net power (approximated by the rate of heat addition) added to the air

flowing through the compressor. Qnet also approximates the change of enthalpy.

Power for 100% efficient isentropic compressor:



P,:=1/o = *CP *Ti* PR k -1 Eq. 5

Pq=looo is an approximated power requirement of an isentropic compression process

based on the measured air temperatures, volumetric flow rates, and pressures. This value









provides a basis for comparison of the relative work performed by each compressor

configuration. By using this value, it is assumed the desired compressor work is only to

increase the pressure ratio of the air and that the temperature of the discharge air is

irrelevant.

Motor controller input power:

MC = V*A Eq. 6


V= measured voltage input to the motor controller

A = measured current input to the motor controller

Effective system efficiency:


77=100%
S1, Eq. 7
MC

For this experimental setup, rsys is the best indicator of the effectiveness of the integration

of a heat pipe in increasing the overall performance of this scroll compressor system.

Availabilty analysis:


AS = cpln e R ln(PR) Eq. 8


AA = lh *((hex, h,)- TAS) Eq. 9

hext= enthalpy at Texit

h,, = enthalpy at T,,

The availbilty between the inlet and exit states of the compressor was calculated for each

case using Eq. 8 and Eq. 9. The values for enthalpy were obtained from tables.5

Availability-based efficiency:









AA
rA = M Eq. 10
MC
The availability-based efficiency, iA, compares the minimum power required to go from

the inlet state to the exit state isothermall compression followed by isentropic heat

addition) and the motor controller input power. This is relevant if the discharge

temperature of the air is important and a higher discharge air temperature is desirablee.' 5

Five test runs were completed for each compressor setup: stock compressor,

compressor with heat pipes and free convection (heat rejection to the atmosphere by free

convection from the heat pipe, Figure 20), and with heat pipes and forced convection

(heat rejection to the atmosphere by forced convection, Figure 23). For each test run, the

ten data points recorded for each variable at each motor speed were averaged to obtain a

single value representing that motor speed and test run. For example, on test run #2 for

the forced convection heat pipe setup, the ten data points for discharge air temperature

taken at a motor speed of 960 rpm were averaged together to obtain a single value of

59.09 C. Next, this value was averaged with the average discharge air temperature

values at 960 rpm from the other 4 test runs to obtain a value of 59.35 C. This final

value is taken to be the standard value for discharge air temperature at a motor speed of

960 rpm for the forced convection heat pipe setup (Table 1). Standard deviations were

also calculated for all variables (Appendix B).










Table 1. Data avera going process sample
Forced Convection Forced Convection
Test 2 960rpm Test run Exit (C)
Data point Exit (C) 1 59.38
1 58.00 2 59.09
2 60.60 3 59.49
3 58.70 4 59.91
4 59.90 5 58.90
5 59.20 Average 59.35


6
7
8
9
10
Average


59.10
58.40
57.70
59.60
59.70
59.09


Uncertainty

The uncertainty for each variable is the combination of three values: the standard

deviation of the ten data points from each test run (Stdrun), the standard deviation of the

average values of each variable from the five test runs of each setup (Stdsetup), and the

instrument error (1). Depending on the variable, the instrument error may consist of

multiple sources of instrument error. For example, the final calculation for system

efficiency includes six sources of instrument error. A sensitivity analysis and the root-

mean-square method were used to calculate the uncertainty due to instrument error. The

final uncertainty (co) calculation also used the root-mean-square method as shown in Eq.

11.


C) = /Stdr 2 Std" 2*2 1


Eq. 11










Experimental Data and Analysis

Motor Speeds

The tests were conducted with motor speeds as close as possible to the ideal speeds

of 900 rpm, 960 rpm, 1020 rpm, and 1080 rpm. Average actual test speeds are listed in

Table 2.

Table 2. Actual test speeds
STOCK
rpm
900.32
960.48
1020.76
1080.78

FREE CONVECTION
rpm
900.27
960.38
1020.21
1079.99

FORCED CONVECTION
rpm
900.41
960.41
1020.28
1080.52


Compressor Component Temperatures

The first parameters of interest are the temperatures of the compressor components:

Scroll 1-2, Fin, Motor, and Pipe 1-3. Table 3 shows the average values from the tests.

The addition of the heat pipe in free convection and forced convection configurations

reduced the temperatures of the fixed scroll wall, fin, and motor. The compressor in the

stock configuration showed a slight temperature gradient between Scroll 1 and Scroll 2;

the cause was hypothesized to be the increasing air temperature as it was compressed and










moved along the flow path. The heat pipe in free convection configuration had the effect

of relatively equalizing the scroll wall temperatures at the thermocouple locations while

the forced convection configuration reversed the stock temperature gradient.

Also of importance was the reduction in operating temperature of the motor. Heat

conducted from the motor to the fixed scroll by way of the compressor housing reducing

the motor operating temperature and potentially increasing the motor efficiency and

longevity.

Table 3. Compressor component temperatures
STOCK Temperatures in C
rpm Scroll 1 Scroll 2 Fin Motor
900 58.30 60.26 53.36 51.46
960 63.06 64.65 56.75 54.97
1020 66.04 68.17 59.65 57.34
1080 70.08 72.66 63.50 61.37

FREE CONVECTION Temperatures in C
rpm Scroll 1 Scroll 2 Fin Motor Pipe 1 Pipe 2 Pipe 3
900 51.44 52.03 46.62 47.36 44.63 44.08 42.55
960 53.72 54.02 48.12 49.63 46.43 46.08 44.32
1020 58.22 58.69 52.34 53.57 50.18 49.75 48.05
1080 61.67 61.92 55.09 56.36 52.62 52.11 50.17

FORCED CONVECTION Temperatures in C
rpm Scroll 1 Scroll 2 Fin Motor Pipe 1 Pipe 2 Pipe 3
900 40.30 39.00 36.17 40.41 28.84 28.86 27.35
960 41.16 40.05 36.92 42.38 28.34 28.39 26.26
1020 44.51 42.39 39.30 45.07 30.06 30.02 28.59
1080 46.30 43.81 40.44 46.57 29.95 29.74 27.86


The heat pipe in free convection configuration was cooler than the scroll wall and

fin and exhibited increased temperatures with increased motor speed. The temperature of

the heat pipe in the forced convection configuration was also cooler than the scroll wall

and fin. The difference between the heat pipe and scroll wall temperatures was greater in

the forced convection configuration compared to the difference in the free convection










configuration, which was due to a greater heat flux. The heat pipe in forced convection

also demonstrated a relatively constant temperature over the range of motor speeds

demonstrating the heat pipe's ability to transfer varying rates of heat while maintaining a

near constant temperature. In this case, the heat transfer rate by forced convection on the

condenser side (fins) of the heat pipe was nearly able to match the increased heat flux

from the evaporator side (fixed scroll). Also, the heat pipe operated nearly isothermally

as expected; thermocouples Pipe 1 and 2 that were placed on the tubes were nearly

identical in all cases and Pipe 3 in the fin being slightly cooler.

Air Discharge Temperature and ATexit-in

The addition of the heat pipe reduced the discharge air temperature up to 25 C

compared to the stock configuration. The intake air temperature, exit (discharge) air

temperature, and their difference, Texit-,,, are listed in Table 4.

Table 4. Air intake and discharge temperatures and ATexit-in
STOCK Temperatures in C
rpm Tin Texit ATexit-in
900 24.99 75.67 50.69
960 24.67 82.10 57.43
1020 24.70 87.20 62.50
1080 24.64 93.56 68.93

FREE CONVECTION Temperatures in C
rpm Tin Texit ATexit-in
900 24.68 67.17 42.48
960 24.73 72.19 47.46
1020 25.39 78.02 52.62
1080 25.20 83.96 58.76

FORCED CONVECTION Temperatures in C
rpm Tin Texit ATexit-in
900 25.09 55.73 30.64
960 24.72 59.35 34.64
1020 25.38 64.35 38.97
1080 24.73 68.05 43.31










Figure 26 shows the reduced AText-m of the heat pipe configurations compared to

stock with the forced convection configuration demonstrating the lowest AText,- for all

motor speeds.


ATexit-in vs. Motor Speed



80.00
stock = 0.0994x 38.548
70.00R2 = 0.9972

60.00
5.0 free = 0.0901x- 38.895
50.00 R2 =0.9974
I-
< 40.00
Forced = 0.0706x 33.014
30.00 --R2= 0.9996

20.00
870 900 930 960 990 1020 1050 1080
motor speed (rpm)

*stock free forced


Figure 26. ATexit-in vs. motor speed.

Mass Flow Rate and Pressure Ratio

The reduced air temperatures due to the reduced temperatures of the fixed scroll

wall directly increased the mass flow rate of air for a given motor speed.

Correspondingly, the pressure ratios achieved by the compressor also increased. Table 5

lists the mass flow rates and the pressure ratios for each case. Figures 27-29 compare the

motor speeds, mass flow rates, and pressure ratios.

The free convection and forced convection heat pipe configurations had increased

mass flow rates and pressure ratios as compared to the stock compressor. The stock and

free convection configurations had nearly identical pressure ratios based on mass flow











rates while the forced convection setup exhibited slightly lower pressure ratios for the

same mass flow rates.

Table 5. Mass flow rates and pressure ratios
STOCK
rpm fii (kg/hr) PR
900 2.18 2.00
960 2.28 2.07
1020 2.38 2.17
1080 2.49 2.25

FREE CONVECTION
rpm fii (kg/hr) PR
900 2.22 2.02
960 2.33 2.12
1020 2.42 2.20
1080 2.53 2.29

FORCED CONVECTION
rpm fii (kg/hr) PR
900 2.26 2.03
960 2.39 2.15
1020 2.49 2.24
1080 2.62 2.36


Mass Flow Rate vs. Motor Speed


J















870 900 930 960 990 1020 1050 1080
motor speed (rpm)
stock Efree forced
stock U free forced


forced = 0.002x + 0.4765
R2 = 0.9986

free = 0.0017x + 0.669
R2 = 0.9985


stock = 0.0017x + 0.6271
R2= 1


Figure 27. Mass flow rate vs. motor speed


















forced = 0.0018x + 0.4315
R2 = 0.997


free = 0.0015x + 0.6926
R2 = 0.9982



stock = 0.0015x + 0.6808
R2 = 0.9937


2.40
2.35
2.30
2.25
2.20
2.15
2.10
2.05
2.00
1.95
8


2.20 2.30 2.40 2.50
mass flow rate (kglhr)


stock = 0.8456x + 0.1508
R2 = 0.9933


free = 0.8564x + 0.1204
R2 = 0.9986


forced = 0.8979x + 0.0027
R2 = 0.9993


2.60 2.70


* stock free forced


Figure 29. Pressure ratio vs. mass flow rate


Pressure Ratio vs. Motor Speed


900 930 960 990 1020 1050 1080
motor speed (rpm)

*stock free forced



Figure 28. Pressure ratio vs. motor speed




Pressure Ratio vs. Mass Flow Rate


70


2.40 -

2.35

2.30

2.25

2.20

2.15

2.10

2.05

2.00

1.95 -
2.10


T










Power Consumption

Table 6 lists the calculated values for MC, Qne, and P=loo%.

Table 6. Calculated power values
STOCK Power in watts
rpm MC Qnet P n=100%
900 93.01 30.78 39.66
960 106.13 36.54 43.77
1020 120.71 41.57 49.32
1080 135.24 47.83 54.11

FREE CONVECTION Power in watts
rpm MC Qnet P n=100%
900 95.37 26.29 41.03
960 109.69 30.88 46.42
1020 124.31 35.58 51.17
1080 140.16 41.50 56.23

FORCED CONVECTION Power in watts
rpm MC Qnet P n=100%
900 98.38 19.33 42.32
960 113.31 23.07 48.64
1020 128.90 27.10 53.93
1080 145.49 31.70 60.63


MC increased for a given motor speed (Figure 30) with the heat pipe installed

because the mass flow rates and pressure ratios were increased. Therefore, to maintain

the same motor speed, more power was required to flow the additional air and compress

the air to an increased pressure ratio. The increased mass flow rates and pressure ratios

also translated into more relative compressor power as shown by the increased values of

P, 1oo (Figure 31). MC and P,= oovwere also compared on a mass flow rate basis

(Figures 32 and 33). This is most applicable to the cases of the stock configuration and

the free convection configuration as they exhibited nearly identical pressure ratios versus

mass flow rates (Figure 29). On a mass flow rate basis, the heat pipe configurations had











lower values ofMC indicating reduced power consumption. The trends for P, loo%


matched those of the pressure ratio trends on a mass flow rate basis. Therefore, the heat


pipe configurations were doing the same relative compressor work with reduced power


consumption (Figure 34).


MC vs. Motor Speed


150.00 fc

140.00
lf
130.00

I 120.00 t

: 110.00

100.00

90.00
870 900 930 960 990 1020 1050 1080
motor speed (rpm)

*stock free forced


d = 0.2615x- 137.45
R2 = 0.9995
=0.2487x -128.91
R2 = 0.9994
=0.2348x- 118.85
R2 = 0.9994


Figure 30. MC vs. motor speed


P n=100% vs. Motor Speed


65.00
fo
60.00 -

55.00 fr

50.00
stt
S45.00

40.00

35.00
870 900 930 960 990 1020 1050 1080
motor speed (rpm)

*stock free forced


ced = 0.1003x 47.952
R2 = 0.9983
e = 0.0841x 34.527
R2 = 0.9995
ck = 0.0813x 33.779
R2 = 0.997


Figure 31. P n=100%vs. motor speed











MC vs. Mass Flow Rate


stock = 136.36x- 204.34
R2= 0.9993

free = 143.97x-224.83
R2= 0.9978

forced = 131.62x- 199.88
R2= 0.9986


150.00

140.00

130.00

I 120.00

110.00

100.00

90.00
2.1


2.20 2.30 2.40 2.50
mass flow rate (kglhr)


forced = 50.55x 72.057
R2=

free = 48.699x- 67.049
R2= 0.9996

stock = 47.179x- 63.353
R2= 0.9968


2.60 2.70


*stock free forced


Figure 33. P n=1oo0vs. mass flow rate


2.20 2.30 2.40 2.50 2.60
mass flow rate (kglhr)

*stock free forced



Figure 32. MC vs. mass flow rate


P n=100o vs. Mass Flow Rate


0


65.00

60.00

- 55.00

50.00
8

45.00

40.00


35.00 I-
2.10


PA'
---


ZZI











MC vs. P r=100%


150.00
/- stock = 2.8848x 20.987
140.00 R2 = 0.9988

130.00 free = 2.9566x 26.638

120.00
S120.0 forced = 2.6034x- 12.251
2 11000 R2 = 0.9984
110.00

100.00

90.00
35.00 40.00 45.00 50.00 55.00 60.00 65.00
P n=100% (watts)

*stock *free forced


Figure 34. MC vs. P = 100%

Using the equations from the linear trendlines established in Figure 34, MC was

calculated for each of the configurations over the range of P=10oo0 tested (Table 7). The

data for the free and forced convection configurations were extrapolated down to a

P,=loo0 of 38 watts even though there were not tested down to that level. Likewise, the

stock and free convection configurations were extrapolated up to a P ,=ooo of 62 watts.

The free convection configuration required a MC of -96.7% relative to the stock

configuration at a P, looo of 38 watts; thus, MC was reduced -3.3%. At a P, loo0 of 62

watts, the free convection configuration showed a reduction of -0.9% in MC relative to

stock. The forced convection configuration showed a reduction ofMC of-2.2% at a

P1=-oo0 of 38 watts and a larger reduction of -5.5% at a Pq,=oo1 of 62 watts. It is

expected the forced convection configuration will further reduce MC relative to the stock

configuration at higher compressor loads.










Table 7. Calculated MC from trendlines for tested range of P,=ioo%0
MC Units in watts Relative to stock
Pn=100% Stock Free Forced Freerelative (%) Forcedrelative (%)
38 88.6 85.7 86.7 96.7 97.8
40 94.4 91.6 91.9 97.1 97.3
42 100.2 97.5 97.1 97.4 96.9
44 105.9 103.5 102.3 97.6 96.6
46 111.7 109.4 107.5 97.9 96.2
48 117.5 115.3 112.7 98.1 95.9
50 123.3 121.2 117.9 98.3 95.7
52 129.0 127.1 123.1 98.5 95.4
54 134.8 133.0 128.3 98.7 95.2
56 140.6 138.9 133.5 98.8 95.0
58 146.3 144.8 138.7 99.0 94.8
60 152.1 150.8 144.0 99.1 94.6
62 157.9 156.7 149.2 99.2 94.5


Effective System Efficiency

The calculated values for the effective system efficiency, sys,, are listed in Table 8

and graphed in Figures 35 and 36. From the experimental data, the heat transfer provided

by the heat pipe increased the effective system efficiency. Comparing sys, on a mass flow

rate basis may give a better indication as to the differences created by the heat transfer

but unfortunately, due to the relatively large uncertainty associated with the rsys

calculation, there is no clear separation between the values from the stock and free

convection or free and forced convection configurations. However, the separations

between the values of the stock and forced convection configurations are mostly well

defined. On either a motor speed or mass flow rate basis, the forced convection

configuration exhibits a higher sys, versus the stock setup. The values of rys, for the free

convection configuration fall between the other two configurations.










Table 8. Effective system efficiencies, i ,
STOCK
rpm rsys (%)
900 42.61
960 41.23
1020 40.84
1080 39.99


FREE CONVECTION
rpm rsys (%)
900 43.00
960 42.30
1020 41.14
1080 40.10

FORCED CONVECTION
rpm rsys (%)
900 42.97
960 42.90
1020 41.81
1080 41.65


44.00
43.50
43.00
42.50
42.00
41.50
41.00
40.50
40.00
39.50
39.00


870 900 930 960 990 1020
motor speed (rpm)

*stock free forced


forced = -0.0084x + 50.629
R2 = 0.8701

free = -0.0164x + 57.91
R2 = 0.9905

stock = -0.0137x + 54.756
R2 = 0.9508


1050 1080


Figure 35. lsys vs. motor speed


vsys VS. Motor Speed


"- _











risys VS. Mass Flow Rate

44.00
43.50
43.50 forced = -4.1595x + 52.489
43.00 R2 = 0.8457
42.50
S42.00 free = -9.4857x + 64.182
S41.50 R = 0.983
41.00
40.50 stock = -7.9696x + 59.761
40.00 R2 = 0.9519
39.50
39.00
2.10 2.30 2.50 2.70
mass flow rate (kglhr)

*stock Efree forced



Figure 36. rsys vs. mass flow rate

Using the equations from the linear trendlines established in Figures 35 and 36, 1 ,

was calculated for each of the configurations over the range of motor speeds and mass

flow rates tested (Tables 9 and 10). When compared over the range of motor speeds

tested, the free convection configuration had a relative 1 of-101.7% compared to the

stock configuration at 900 rpm. Hence, the free convection configuration showed an

improvement of-1.7%. This improvement decreased to -0.6% at 1080 rpm. The forced

convection configuration showed an improvement of-1.5% at 900 rpm which increased

to -4.0% at 1080 rpm.

Table 9. Calculated ys, from trendlines for tested range of motor speeds
rsys Units in % Relative to stock
rpm Stock Free Forced Freerelative Forcedrelative
900 42.4 43.2 43.1 101.7 101.5
930 42.0 42.7 42.8 101.5 101.9
960 41.6 42.2 42.6 101.4 102.3
990 41.2 41.7 42.3 101.2 102.7
1020 40.8 41.2 42.1 101.0 103.1
1050 40.4 40.7 41.8 100.8 103.6
1080 40.0 40.2 41.6 100.6 104.0










For comparison on a mass flow rate basis, the data for the free and forced

convection configurations was extrapolated down to a mass flow rate of 2.18 kg/hr.

Likewise, the stock and free convection configurations were extrapolated up to 2.62

kg/hr. From Table 10, the free convection configuration showed a relative increase of

-2.6% at 2.18 kg/hr flowrate which reduced to -1.2% at 2.62 kg/hr flowrate. The forced

convection configuration showed an improvement of -2.4% at 2.18 kg/hr flowrate which

increased to -7.0% at 2.62 kg/hr flowrate.

Table 10. Calculated sys from trendlines for tested range of mass flow rates
rsvs Units in % Relative to stock
_ih (kg/hr) Stock Free Forced Freerelative Forcedrelative
2.18 42.4 43.5 43.4 102.6 102.4
2.22 42.1 43.1 43.3 102.5 102.8
2.26 41.7 42.7 43.1 102.4 103.2
2.30 41.4 42.4 42.9 102.3 103.6
2.34 41.1 42.0 42.8 102.1 104.0
2.38 40.8 41.6 42.6 102.0 104.4
2.42 40.5 41.2 42.4 101.9 104.8
2.46 40.2 40.8 42.3 101.7 105.2
2.50 39.8 40.5 42.1 101.6 105.7
2.54 39.5 40.1 41.9 101.4 106.1
2.58 39.2 39.7 41.8 101.3 106.5
2.62 38.9 39.3 41.6 101.2 107.0


Availability-Based Efficiency

The calculated values for rlA are listed in Table 11 and graphed in Figures 37 and

38. The calculated rlA was lower for all motor speeds for the heat pipe configurations

compared to the stock configuration. This was also the general trend for the range of

mass flow rates covered by all three configurations. This was due to the lower discharge

temperatures of the air from the heat pipe configurations resulting in a lower change in

enthalpy.











Table 11. Calculated availabilty-based efficiency


STOCK
rpm AA (watts) MC (watts) nA (%)
900 38.24 93.01 41.11
960 42.36 106.13 39.92
1020 47.53 120.71 39.38
1080 52.71 135.24 38.98


FREE CONVECTION
rpm AA (watts) MC (watts) qA (%)
900 38.59 95.37 40.46
960 43.69 109.69 39.83
1020 48.73 124.31 39.20
1080 53.67 140.16 38.29


FORCED CONVECTION
rpm AA (watts) MC (watts) qA (%)
900 39.13 98.38 39.78
960 44.55 113.31 39.32
1020 49.84 128.90 38.67
1080 55.43 145.49 38.09


qA vs. Motor Speed


900 930 960 990 1020
motor speed (rpm)

*stock free forced


stock = -0.0115x + 51.267
R2 = 0.9345

free = -0.0119x + 51.232
R2 = 0.9906

forced = -0.0095x + 48.374
R2 = 0.996


1050 1080


Figure 37. TA vs. motor speed


42.00
41.50
41.00
40.50
40.00
39.50
39.00
38.50
38.00
37.50
37.00
8


70












qA vs. Mass Flow Rate


42.00
41.50
41.00
40.50
40.00
39.50
39.00
38.50
38.00
37.50
37.00


2.10 2.20 2.30 2.40 2.50
mass flow rate (kglhr)
*stock free forced


stock = -6.6984x + 55.473
R2 = 0.9354


free = -6.8926x + 55.83
R2 = 0.9898

forced = -4.7746x + 50.624
R2 = 0.9921


2.60 2.70


Figure 38. lA vs. mass flow rate

Using the equations from the linear trendlines established in Figures 37 and 38, lA

was calculated for each of the configurations over the range of motor speeds and mass

flow rates tested (Tables 12 and 13). The rA for the free convection configuration

relative to stock was -1.0% lower at a motor speed of 900 rpm and -1.2% lower at a

motor speed of 1080 rpm. The rlA for the forced convection configuration relative to

stock was -2.7% lower at 900 rpm but the difference reduced to -1.9% at 1080 rpm.

Table 12. Calculated TA from trendlines for tested range of motor speeds
hA Units in % Relative to stock
rpm Stock Free Forced Freerelative Forcedrelative
900 40.9 40.5 39.8 99.0 97.3
930 40.6 40.2 39.5 99.0 97.5
960 40.2 39.8 39.3 99.0 97.6
990 39.9 39.5 39.0 98.9 97.7
1020 39.5 39.1 38.7 98.9 97.8
1050 39.2 38.7 38.4 98.8 98.0
1080 38.8 38.4 38.1 98.8 98.1


T










Comparision on a mass flow rate basis showed the free convection configuration

having a rlA -0.2%-0.4% lower the than stock configuration over the extrapolated range.

The forced convection configuration had a rlA -1.6% lower than stock at a 2.18 kg/hr

flow rate. However, it had a rlA -0.5% higher when compared at a 2.62 kg/hr flow rate.

It is expected that the forced convection configuration will have an increased rlA relative

to stock at higher than tested flow rates.

Table 13. Calculated rlA from trendlines for tested range of mass flow rates
hA Units in % Relative to stock
Mih (kg/hr) Stock Free Forced Freerelative Forcedrelative
2.18 40.9 40.8 40.2 99.8 98.4
2.22 40.6 40.5 40.0 99.8 98.6
2.26 40.3 40.3 39.8 99.8 98.8
2.30 40.1 40.0 39.6 99.8 98.9
2.34 39.8 39.7 39.5 99.8 99.1
2.38 39.5 39.4 39.3 99.7 99.3
2.42 39.3 39.1 39.1 99.7 99.5
2.46 39.0 38.9 38.9 99.7 99.7
2.50 38.7 38.6 38.7 99.7 99.9
2.54 38.5 38.3 38.5 99.6 100.1
2.58 38.2 38.0 38.3 99.6 100.3
2.62 37.9 37.8 38.1 99.6 100.5


Approximated Compressor Efficiency

It was possible to approximate compressor efficiency, rap, comparing shaft power

to isentropic compression power (P =100oo) from the experimental data using the

specifications of the compressor. The forced convection configuration at 1080 rpm had

the following average values: MC = 145.49 watts, volumetric flow rate = 36.01 LPM,

and Pdischarge,gage = 20.44 psi (141 kPa) (Appendix B). This closely matches the

specifications for the compressor at 1000 rpm of: volumetric flow rate = 36 LPM,

Pdischarge,gage = 20 psi (138 kPa), and shaft power = 109 watts (Appendix A). From this

information, the approximate percentage of power converted from MC to compressor










shaft power is 75%. Assuming this percentage remains constant for all cases, an

approximated shaft power for each case can be calculated from Eq. 12. This assumption

was deemed reasonable because of the limited range of motor speeds tested. rqp was

calculated from Eq. 13. Table 14 lists the approximated shaft power, SPpprox, and rqs for

each case. Figures 39 and 40 show the trends of rql are the same as ,

Approximated shaft power:


SP = MC 0.75

Approximated compressor efficiency:


sp 7100
approx

Table 14. Appoximated compressor shaft power and rlsp
STOCK Power in watts
rpm MC SPapprox Isp(%)
900 93.01 69.8 56.9
960 106.13 79.6 55.0
1020 120.71 90.5 54.5
1080 135.24 101.4 53.3

FREE CONVECTION Power in watts
rpm MC SPapprox sp (%)
900 95.37 71.5 57.4
960 109.69 82.3 56.4
1020 124.31 93.2 54.9
1080 140.16 105.1 53.5

FORCED CONVECTION Power in watts
rpm MC SPapprox spl(%)
900 98.38 73.8 57.4
960 113.31 85.0 57.2
1020 128.90 96.7 55.8
1080 145.49 109.1 55.6


Eq. 12


Eq. 13







46




rsp vs. Motor Speed


59.00 -

58.00

57.00

56.00

S55.00

54.00

53.00

52.00
870


Figure 39. fsp vs. motor speed



qsp vs. Mass Flow Rate


2.20 2.30 2.40 2.50 2.60
mass flow rate (kglhr)

*stock free forced


Figure 40. rsp vs. mass flow rate


900 930 960 990 1020 1050 1080
motor speed (rpm)
*stock free forced


59.00 -

58.00

57.00

56.00

55.00

54.00

53.00

52.00 -
2.10













CHAPTER 5
AIR COMPRESSION MODEL

Modeling Process

A spreadsheet was created in Excel to generate a simple model of the air during

compression to determine compressor power, air temperature, and the heat removal rate

from the air. Each of the three setups and four motor speeds (twelve cases total) were

modeled, as was an isentropic compression of air using the same inputs. The model was

based on the known internal volume ratio of the compressor (1.75:1). The compression

process was modeled as 27 stages. The first stage represented the mass of air that just

entered the compressor and had not yet been compressed. Each of the next 25 stages

compressed the air mass in increments of 0.03 with respect to the overall volume ratio

(i.e. from 1.00:1 to 1.75:1). The final stage represented the air leaving the final gas

pocket and entering the discharge reservoir.

Each case was modeled using the mass flow rate, air intake temperature, pressure

ratio, scroll wall temperature and discharge air temperature from the experimental data.

Each stage of the model makes the following calculations in the order listed:

Volume ratio:



V
VR = /72 Eq. 14


Pressure ratio 1:


PR= P = V Eq. 15
[P1 V2









k = ratio of constant specific heats =1.4 for air

Ideal compressor work:


W,(J)= m*Cp*T* PR k -1 Eq. 16


m = mass of air in the gas pocket
Ideal compression temperature increase:


AT= T,* PRk -1 Eq. 17


Intermediate temperature:

T2 = T + AT Eq. 18

Air temperature difference from scroll wall temperature:

AT scroll = Tall T2 Eq. 19

Twa11 = scroll wall temperature

Heat removed:


Q moved (J) = H** A ** ATcrozA Eq. 20

H = model heat transfer coefficient (J/m2*K)

a = relative heat transfer area coefficient

A, = initial heat transfer surface area

A, was the approximate surface area of the gas pocket after it had been completely

sealed from the intake port and uncompressed. The heat transfer surface area at the

compressor discharge was approximated as being half ofA,. Therefore, the relative heat

transfer area, a, was used to adjust the heat transfer surface area relative to A, to account









for the decreased gas pocket volume, a = 1 for the first stage of compression and reduced

linearly in increments of 0.02 for each stage resulting in a = 0.5 for the discharge stage.

Final temperature:


T3 moved +T, Eq. 21
m*Cp

Pressure ratio 2:


PR, = P = Eq. 22


Overall pressure ratio:


PRtage = PR1 *PR2 PRstage,prevous Eq. 23


PRstage,previous = PRstage from the proceeding stage

Ts becomes T1 of the following stage and the loop of calculations is performed until

all 27 stages are completed. The first stage of the model assumes: PR1 = 1, T2 = T,, from

the experimental data, Ts = approximated temperature, PR2 = 1, and PRtage= 1. T3 was

approximated because of the hypothesis that the difference in mass flow rates between

each setup for a given motor speed was due to the difference in initial heating of the air.

Hence, the heat pipe setups had higher mass flow rates relative to stock due to the air

being heated less and remaining denser. Ts for the stock setup was approximated as the

mean of T, and Twai. Based on the differences in mass flow rates, the air densities and

temperatures for the first stage of the heat pipe setups were calculated (Table 15). Once

Ts was determined, Qremoved for the first stage could be calculated from the Eq. 24, which

is only used for the first stage (remaining stages use Eq. 20).









Qremoved,1 = *C *(T3 -T2) Eq. 24

PR from the experimental data was entered into the final step and used to calculate

PR1,27 from Eq. 25. This made the overall model pressure ratio match the experimental

data. It was assumed that the heat transfer in the final step did not change the pressure

ratio, as the air was no longer confined to a finite size volume.

PR
PR1,27 = Eq. 25
PRtage, previous

The model heat transfer coefficient H was adjusted until the final discharge air

temperature in the model matched the experimental data. Once this value was finalized,

the total compressor power, heat removal rate, effective efficiency, and heat transfer

coefficient he were calculated from the equations below. The motor revolutions per

second (rps) from the experimental data were used for each case and the multiplier (2)

accounts for two gas pockets entering and exiting the compressor each revolution.

Total compressor power:


TCP(W)= WC rps 2 Eq. 26

Total heat removal rate:


Qemoved,total (W)= Q removed* r* 2 Eq. 27

Heat transfer coefficient:


he(W/m2*K)=H*rps*2 Eq. 28

The total compressor power, TCP, represents the power used directly in the process

of compressing the air from when the air enters the compressor to being discharged.

Qremoved,total is the rate of heat removal for this process.









Furthermore, the model calculated the values Qnet,, and P, =Ioo0o,m corresponding to


Qnet and P, o10o0 from the experimental data analysis to verify the accuracy of the model.

Net rate of heat addition:


Qnet,m = TCP Qremoved,total Eq. 29

P1 -o00o was calculated using the model with H = 0 and the first stage having the

values of PR1 = 1, T, = T,,.

Additional assumptions were made in the model. One assumption was a constant

scroll wall temperature, Twa,,, along the flow path. Thermocouples Scroll 1 and 2 from

the experimental tests showed a difference of 1-2 C (Table 3) which was within

uncertainty (Appendix B); therefore, using an average value of Scrolls 1 and 2 for Twazi

was deemed a suitable approximation. The next assumption was that the temperature of

the orbiting scroll wall equilibrated to the fixed scroll wall temperature at steady state.

This assumption seems valid based on the difference between Twail and the discharge air

temperature (ATexa-wall) being within 1-2 C for each setup at a given motor speed (Table

16). Lastly, it was assumed that there was no leakage between air pockets.2

Model Results and Analysis

The air compression model was used to calculate temperature profiles, power

distribution and heat removal rates that could not be directly calculated from the

experimental data. The model results are suitable for comparison to each other; however,

the model results are not truly accurate as the discharge air temperature values from the

experimental data were a required input.









Temperature Profiles

The model plotted the air temperature during compression against an isentropic

case, based on the same air mass flow rate and pressure ratio, and also the scroll wall

temperature. The two cases of the stock configuration at 900 rpm and the forced

convection heat pipe configuration at 1080 rpm are plotted in Figures 41 and 42. Graphs

for all the configurations are in Appendix D.

The model predicted the temperature of the air compressed by the stock

configuration being higher than an isentropic compression for over half of the

compression process (Figure 41). The scroll wall temperature had a direct effect on the

air temperature during compression. The slope of the line representing the temperature

of the air compressed by the scroll compressor (Tactal) was steeper than the isentropic

compression (Tientropi) until the air temperature reached the temperature of the scroll

wall; hence, the air was being heated by the scroll wall. Once the temperature of the air

exceeded the temperature of the scroll wall, the scroll wall extracted heat from the air and

the slope of the line for the scroll compressor air was less than the isentropic compression

process. The curve of Tactual demonstrates the path dependent nature of non-

adiabatic/non-isentropic compression and why the general isentropic efficiency

calculation based on the inlet and exit conditions is not representative for this type of

compression process.

In comparison to the stock configuration, the forced convection heat pipe

configuration (Figure 42) shows the cooling effect of the scroll wall more dramatically.

Tactual exceeded Tsentropic for half the portion of the previous case. The cooling effect of

the scroll wall then kept the scroll compressor air significantly cooler than an isentropic

process. In all cases, the model predicted the air was compressed nearly isothermally











after a point. Also, note the jump in the temperature profile in the last step for all cases.

This was due to the pressure of the air leaving the compressor equilibrating with the


higher pressure of the discharge reservoir. This also meant the compressor was required


to do more work to compress the air during the discharge stage.


Table 15 lists the approximated air temperatures and densities for the first stage in


the model for each case. Table 16 lists the results AText-waii


Model Predicted Temperature Profile-Stock 900 rpm


110
100
0 90
80
- 70
2 60
. 50
40
30
20


1 1.1 1.2 1.3 1.4 1.5 1.6 1.7
volume ratio

-- Tactual -- Tisentropic Twall


Figure 41. Model predicted temperature profile for stock configuration at 900 rpm


Model Predicted Temperature Profile-Forced 1080 rpm


S100

80

S60

E 40

20


1 1.1 1.2 1.3 1.4 1.5 1.6 1.7
volume ratio

---Tactual ---Tisentropic Twall


Figure 42. Model predicted temperature profile for forced convection configuration at
1080 rpm










Approximated model first stage air temperatures and densities


STOCK
rpm T (C) p (kg/m3)
900 41.64 1.143
960 43.87 1.135
1020 45.37 1.130
1080 47.36 1.123

FREE CONVECTION
rpm T (C) P (kg/m3)
900 35.74 1.165
960 36.92 1.161
1020 40.18 1.149
1080 41.75 1.143

FORCED CONVECTION
rpm T (C) P (kg/m3)
900 29.82 1.188
960 29.77 1.188
1020 31.46 1.181
1080 30.79 1.184

Table 16. ATexit-wall values
STOCK Temperatures in C
rpm Twall Texit ATexit-wall
900 58.85 75.67 16.82
960 63.85 82.10 18.25
1020 66.85 87.20 20.35
1080 70.85 93.56 22.71

FREE CONVECTION Temperatures in C
rpm Twall Texit ATexit-wall
900 51.85 67.17 15.32
960 53.85 72.19 18.34
1020 57.85 78.02 20.17
1080 61.85 83.96 22.11

FORCED CONVECTION Temperatures in C
rpm Twall Texit ATexit-wall
900 39.85 55.73 15.88
960 40.85 59.35 18.50
1020 42.85 64.35 21.50
1080 43.85 68.05 24.20


Table 15.









Experimental Data and Model Results Comparison

Table 17 lists the results for ATexit-a, Qn,,, and P, =100o from the experimental test

data and the model results. As expected, ATex ,t- of the test data and model are nearly

identical as a result of the model being adjusted until the discharge air temperature, and

therefore ATex,,tn, nearly matched. The near identical numbers for Qnet and P,= oo0 show

that the model was able to accurately calculate compressor power.

Power Distribution and Heat Removal Rate

The calculations performed by the model allowed for a power distribution

calculation using the following equations. The calculated values are listed in Table 18.


MC Q net + Q removed,total + MElosses + Motorlosses Eq. 30
MElos = Sppox TCP Eq. 31

Motorlosses accounts for the combined electrical and mechanical losses of the motor

controller and motor. MEiosses accounts for the mechanical losses in the compressor.

These losses included fricitional losses and flow losses due to moving the air.

The trends for Qremovedtotal (Figures 43 and 44) are based on the method of heat

transfer from the compressor to the environment. The heat removal rate for the stock and

free convection heat pipe configurations were relatively constant over the range of motor

speeds tested with this being attributed to their heat transfer rates being limited to free

convection and radiation. Therefore, the heat transfer rate was directly limited to the

difference in temperatures between the heat transfer surfaces and the environment. The

free convection heat pipe configuration demonstrated a higher heat removal rate

compared to stock due to the greater surface area. In contrast, the forced convection heat

pipe setup was able to increase the heat removal rate with the increased motor speeds.









The free convection and forced convection configurations exhibited higher

Motoriosses and MElosses for a given motor speed (Figures 45 and 46). The increased

Motoriosses for the heat pipe configurations may be attributed to the higher power

consumption for the same motor speed and the inefficiencies of the motor and motor

controller. The increased MElosses for the heat pipe configurations is attributed to the

greater mass flow rates and pressure ratios leading to higher flow losses. However, when

comparing Motoriosses and MElosses on a mass flow basis (Figures 47 and 48), the heat pipe

configurations exhibited lower losses due to their ability to flow the same mass flow rate

as the stock configuration at a lower motor speed. This provides a good explanation as to

one of the methods that the heat pipe configurations improved effective system

efficiencies calculated from the experimental data.

The values of the approximate heat transfer coefficient he calculated by the model

are listed in Table 19. The heat pipe configurations had higher values of he compared to

the stock configuration for a given motor speed as could be predicted based on the lower

AText,,, observed experimentally and the model predicted Qremoved,total *







57


Table 17. Comparison of experimental data and model results
STOCK
ATexit-,n (C) Qnet (watts) P =100o(watts)
rpm Test Model Test Model Test Model
900 50.69 50.69 30.78 30.78 39.66 39.59
960 57.43 57.43 36.54 36.54 43.77 43.70
1020 62.50 62.50 41.57 41.56 49.32 49.24
1080 68.93 68.93 47.83 47.83 54.11 54.02


FREE CONVECTION
ATexit-,n (C) Qnet (watts) Pn =oo0 (watts)
rpm Test Model Test Model Test Model
900 42.48 42.48 26.29 26.29 41.03 40.96
960 47.46 47.46 30.88 30.88 46.42 46.34
1020 52.62 52.62 35.58 35.58 51.17 51.08
1080 58.76 58.76 41.50 41.50 56.23 56.13


FORCED CONVECTION
ATexit-,n (C) Qnet (watts) Pn =10/(watts)
rpm Test Model Test Model Test Model
900 30.64 30.64 19.33 19.33 42.32 42.22
960 34.64 34.64 23.07 23.07 48.64 48.52
1020 38.97 38.97 27.10 27.10 53.93 53.83
1080 43.31 43.31 31.70 31.69 60.63 60.53


Qdotremoved,total vs. Motor Speed


40.00

7 35.00

. 30.00

- 25.00

S20.00
15
O 15.00

10.00


forced = 0.0356x 1.0223
R2 = 0.9875

free = 0.0113x + 13.114
R2 = 0.7889

stock = 0.0011x + 16.274
R2= 0.0217


870 900 930 960 990 1020 1050 1080
motor speed (rpm)

*stock free forced


Figure 43. Model predicted Qdotremoved, total vs. motor speeds







58


Table 18. Model predicted power distribution
STOCK Power in watts
rpm MC Qnet Qremovedtotal MElosses Motorlosses
900 93.01 30.78 17.49 21.48 23.25
960 106.13 36.54 16.61 26.44 26.53
1020 120.71 41.57 17.96 31.00 30.18
1080 135.24 47.83 17.26 36.34 33.81


FREE CONVECTION Power in watts
rpm MC Qnet removed,total MElosses Motorlosses
900 95.37 26.29 22.98 22.27 23.84
960 109.69 30.88 24.28 27.11 27.42
1020 124.31 35.58 25.18 32.48 31.08
1080 140.16 41.50 24.94 38.67 35.04


FORCED CONVECTION Power in watts
rpm MC Qnet Qremovedtotal MElosses Motorlosses
900 98.38 19.33 30.74 23.71 24.59
960 113.31 23.07 33.58 28.33 28.33
1020 128.90 27.10 35.14 34.42 32.22
1080 145.49 31.70 37.34 40.09 36.37


Qdotremoved,total VS. Mass Flow Rate


40.00

35.00

30.00

25.00

20.00

15.00

10.00


forced = 17.945x- 9.6179
R2 = 0.9914


free = 6.5757x + 8.715
R2 = 0.79


stock = 0.609x + 15.909
R2 = 0.021


2.10 2.20 2.30 2.40 2.50
mass flow rate (kglhr)


2.60 2.70


*stock free forced


Figure 44. Model predicted Qdotremoved, total vs. mass flow rates













Motoriosses vs. Motor Speed


!70


45.00


40.00


S35.00


S30.00
LU

25.00


20.00 -
870


38.00

36.00

34.00

^ 32.00

30.00

o 28.00

S26.00

24.00


900 930 960 990 1020 1050 1080
motor speed (rpm)

stock m free forced


Figure 46. Model predicted MElosses vs. motor speed


forced = 0.0654x 34.362
R2 = 0.9995


free = 0.0622x 32.227
R2 = 0.9994


stock = 0.0587x 29.712
R2 = 0.9994





900 930 960 990 1020 1050 1080

motor speed (rpm)

stock m free forced




Figure 45. Model predicted Motoriosses vs. motor speed



MEIosses vs. Motor Speed




forced = 0.092x 59.502
_R2 = 0.9971


free = 0.0911x 60.117
R2 = 0.9968

stock = 0.0817x 52.086
R2 = 0.9991


22.00
8








60




Motorlosses vs. Mass Flow Rate


38.00

36.00

34.00

S32.00

S30.00

. 28.00
o
26.00

24.00

22.00
2.


stock = 34.089x- 51.085
R2= 0.9993



free = 35.992x- 56.208
R2 = 0.9978



forced = 32.904x 49.969
R2 = 0.9986


2.60 2.70


Figure 47. Model predicted Motoriosses vs. mass flow rate



MEiosses vs. Mass Flow Rate


2.30 2.40 2.50
mass flow rate (kglhr)

*stock free forced


stock = 47.429x- 81.83
R2= 0.9991



free = 52.733x- 95.219
R2 = 0.9944


forced = 46.266x- 81.34
R2 = 0.9938


2.60 2.70


Figure 48. Model predicted MElosses vs. mass flow rate


2.20 2.30 2.40 2.50
mass flow rate (kglhr)

*stock free forced


10


45.00


40.00


S35.00


S30.00
w


25.00


20.00
2.1


0










Table 19. Model calculated approximate heat transfer coefficient he
STOCK
rpm hc (W/m2*K)
900 10
960 11
1020 14
1080 15

FREE CONVECTION
rpm h, (W/m2*K)
900 12
960 14
1020 17
1080 19

FORCED CONVECTION
rpm hO (W/m2*K)
900 13
960 17
1020 18
1080 20














CHAPTER 6
CONCLUSIONS

The purpose of this investigation was to determine the performance effects of heat

transfer from a scroll compressor through the use of a heat pipe. It was tested in three

configurations: stock, integrated with a heat pipe rejecting heat by free convection, and

integrated with a heat pipe rejecting heat by forced convection. Each setup was tested

over a range of motor speeds and a model was developed in Excel to provide a means to

further analyze the experimental data.

The heat transfer allowed by integration of the heat pipe decreased the temperatures

of the fixed scroll wall, fixed scroll, motor, and compressor discharge air. The forced

convection configuration exhibited a scroll wall temperature up to -29 OC cooler and a

ATexiti up to -26 C lower than stock over the tested range of motor speeds. These

reductions in temperatures led increased mass flow rates of over 5% at a -5% higher

pressure ratio.

The power input to the motor controller to compress the air to a specific pressure

ratio at a specific mass flow rate was reduced by -5.5% by the forced convection

configuration relative to stock as calculated using the extrapolated P, =10o0 data. The

reduction in power required resulted in the calculated effective system efficiency and

approximated compressor efficiency to be up to -7% higher for the forced convection

configuration relative to stock. Furthermore, it is predicted that the forced convection

configuration will show more gains at higher compressor loads than tested. These results

are applicable when the cooler discharge temperature of the air is acceptable.









Efficiency calculated based on an availabity analysis, rA, resulted in the free

convection heat pipe configuration rlA being lower than the stock configuration in all

situations. The forced convection configuration exhibited a lower rlA relative to stock at

the lower tested mass flow rates but had a higher rlA at the greater mass flow rates.

These effects may be attributed to the rate of heat removal from the compressor that

was predicted by the model. The model showed the stock configuration having a near

constant rate of heat removal over the range of motor speeds tested. The forced

convection configuration showed a trend of higher heat removal rates that also increased

with motor speed; the heat removal rate was -175% compared to stock at the lowest

motor speed and increased to -220% at the highest motor speed.

Recommendations for the extension of this work include optimizing the

instrumentation, integration of the heat pipe and increasing the compressor load. A high

level of instrument uncertainty led to some of the data calculations and comparisons

being inconclusive and a large source of that error may be attributed to the temperature

measurements. It is recommended that T-type thermocouples be used as opposed to K-

type thermocouples for improved resolution. Furthermore, higher accuracy in pressure

measurement is desirable; however, if an electronic transducer were to be utilized, it

would require a very high sample rate to measure the oscillations inherent in the scroll

compressor output. A better integration of the heat pipe with the scroll compressor is

also desirable to increase the heat transfer surface area between the two. A scroll

compressor designed specifically with heat pipes integrated would be the optimal

situation. Testing at higher compressor loads is desirable to investigate potentially

greater effects than those observed. From the trends established in the experimentation, it









is predicted that a forced convection heat pipe configuration will exhibit the same effects

observed relative to stock at greater magnitudes.

Lastly, the performance of this compressor may not be up to the standard level for

the reason that the experimentation required disassembly and reassembly of the

compressor assembly. Therefore, it is unknown whether the scrolls were reassembled to

the proper clearances.

This research shows a definitive performance advantage associated with heat

transfer from scroll compressors through the use of a heat pipe. Heat transfer from the

compressor increased the mass flow rate of the compressor for a given motor speed,

which is analogous to increasing the energy density of a battery or the specific power

output of an internal combustion engine. Furthermore, power input to the compressor

system was reduced for a given mass flow rate and pressure ratio. The actual benefit of

the addition of a heat pipe to scroll compressors will be dependent on the overall system

to which the compressor will be integrated. System constraints include cost, packaging,

and the method of heat transfer from the heat pipe. The optimal situation would be a

system with another fluid stream requiring heating. Therefore, the heat pipe can extract

heat from the fluid being compressed by the scroll compressor and reject the heat to the

additional fluid stream requiring heat.













APPENDIX A
EXPERIMENTAL COMPONENTS

Air Squared, Inc.
3001 Industrial Lane #3
Broomfield, CO 80020
Air Squared P16H30N2.50 Prototype Compressor


I-
. ,


< W-1"1


lid- L


P16H30N2.50 1000 109 20 42 42 40 38 36
(Prototype) 2000 187 20 50 78 77 75 73


3000 270 20


P16H30N2.50 10.0*


- 115 113 110 108


4.9 6.3 4-Jan


Scroll wall height: 30mm
Scroll wall thickness: 4 mm
Distance between scroll walls: 12mm









Three-Phase Brushless DC motor:
Poly-Scientific, part # BN34-35AF-02CH

Motor Controller:
Poly-Scientific, part # BDO-Q2-50-18
20-50 VDC
6-7 A

Power Supply:
Samlex 120 V AC-to-DC
Adjustable 0-30V, 5A

Heat Pipe Technology, Inc.
4340 NE 49th Avenue
Gainesville, FL 32609
Custom U-shaped Heat Pipe:
12.7 mm diameter copper tubes
25 mm x 63.5 mm aluminum fins

AOS Heat Sink Compound:
part # 52022JS

Y.S. Tech computer fans:
DC 12V, 2.64 W, part # NFD1260157B-1A

Volumetric Flow Meter:
McMillian, Model # S-110-12
0-100 LPM flow range
0.1 LPM resolution
0.40% full-scale error at 20 LPM
0.00% full-scale error at 50 LPM

Compressor Discharge Pressure Gauge:
Autometer, Instr. No. 2650-566
0-35 psi (0-241 kPa) range
0.5 psi (3.45 kPa) resolution
2% full-scale error

Ambient Air Pressure:
Oregon Scientific Indoor Weather Station
0.01 in. Hg resolution

Thermocouples:
Omega, part # 5SRTC-GG-K-30-72
K-type
30 AWG gage, 0.25 mm Dia.wire










Data Acquisition: lotech DBK52 14-channel thermocouple module
lotech DBK24 24-channel isolated digital input module
Daqview software
Dell laptop computer

Fluke 79111 Digital Multi-meters
0.01 V, A, Hz resolution

9.5 mm I.D. clear vinyl tubing

9.5 mm I.D. 50psi fuel line

Brass 14 NPT-to-hose barb fittings

Mueller Industries, Inc. ball valve
Part# R850, 150 WSP, 600 WOG















APPENDIX B
EXPERIMENTAL RESULTS AND UNCERTAINTY1


1 Please refer to Chapter 4



















STOCK

. .: i T ir T ii t,-, I. : f1 : ,., I I -r. .: ',1 -1 : ) 11 I'l P F 1 -: -l ir : : :




stock 4 25.27 76.55 93.70 60.02 29.96 6.02E-04 14.85 1.99 31.00 41.95 58.65 61.04 54.02 51.83
stock 24.97 75.94 93.15 60.08 29.95 6.05E-04 14.98 2.00 30.95 42.49 58.92 60.55 54.63 51.61
average 24.99 75.67 93.01 60.02 30.02 6.05E-04 14.97 2.00 30.78 42.61 58.30 60.26 53.36 51.46
std dev 0.33 0.74 0.45 0.03 0.10 2.29E-06 0.07 0.00 0 27 0.46 0.46 1.01 0 93 0.50
S ... : Tir. T i 1-. I ,: fl 1 .:. l 1 i'.1 r,.l.:.1 .11|.' .:i FF .I.:.,. i|,,, '" :.: .:.ll 1 :.: .:. ll .: ,ir. ir.:.l.:.,
: 1.:..: I -.4 1I 4 I i, -" 4 '4 1:,4 1 4 i .- I-,.,4 I !- 11 4 4 1 i? : :- _f f ; 1-"
'II '-
: 1.:..:1 -. : ; ; I,)- 4 1.4 :1- "1 44 ; .-E ."14 If I, ": 1 -' 4 1 -" 4 I 4 I -:I : ,., : 4 I:i ;

: 1.:..:1 4 .4 ( 4 :1 I i:,i i : .:,n, 1 4 : i ":E .-,,4 If r i:, ,1, :( :] 4 1 1 : i I': 4 4 z -- 4 4 ;
:|... I -- I : ,:,:, I,:, 4 :- : : E 4 I 1: I'- : 41: ~ 4 : 4 :, :, -

average 24.67 82.10 106.13 64.03 31.43 6.34E-04 16.01 2.07 36.54 41.23 63.06 64.65 56.75 54.97
stddev 0.57 0.72 0.73 0.02 0.07 1.1 E-06 0.12 0.01 0 21 0.31 0.98 1.47 1 56 1.16
S .- 5.j.- : Tir., T i1 1.- I.: fl1.:. P_ 1)i' 1 ir.l.:.| '1 .-1:: l .11 I' .:i', FF I.:.,,, i ), ." '" :.: .:.-11 1 : :. :.ll: fir, ir,.:..:..
: :: I -4 ;: :; : l "': ": :4 4i ri -I.'4 I 4- 4 1 i, 4 -f ii : : 1:, 4 -
: .I- 'I 411 .4 1
:1.:.:1 -i 4 lI, --, ,4:, : 1 :,', "" I- ( f -E-',4 l *: Cl, 4,1 4ff i 4-1 4 ,:- : :. :. -4

:I:: -4 : I 1 r.i'i4 I I I 41 4 1 4- r I-- -1 4i -4
: I.:..:I I 4 14: l l4 "" C E :- i I I- f "1 1 '1 I1 4 i- -4- 4

average 24.70 87.20 120.71 68.05 32.85 6.62E-04 17.62 2.17 41.57 40.84 66.04 68.17 59.65 57.34
std dev 0.24 0.75 0.70 0.03 0.05 1.02E-06 0.20 0.01 0.45 0.40 0.67 0.72 0.61 0.93
5 .-i .- : Tir, T i1 i-. I,-: fl.:. 1_ i' l r.l.:.l i l .l :' 11 I .:l i FF n. .:. ., ,,,, ,'" ,1 :.:-.:.l 1 f : .: 1 :ll: hir. i r.:..:.

: I 1.:..: -: ;4 --4 1 4 41 -4 4 :, :E."4 I 1 4 4 ,I1 1 4 4 1 : 4 i
:I.:..:I : : :, --- I"4 I- 1:1 4 :, 11 i E -," 4 I :-I -f 4 -- 41 ': ( 1" I: I4- 1 l
: .:..: 1 -' I -. 4 r I l :- : J :- I:!

:1:.: 1 -4 : ,4 4- i -' :,4 "4 1 lE.,I 4 I 4- 4; 4 1:. -I -I -4-: 4 -: : :

average 24.64 93.56 135.24 72.05 34.27 6.91 E-04 18.81 2.25 47.83 39.99 70.08 72.66 63.50 61.37
std dev 0.74 0.99 1.00 0.01 0.05 1.95E-06 0.09 0.01 0.54 0.33 0.99 1.13 1.10 1.36


1020 i


















FREE
averages Tin I Teit MC hz Iflow (LPFM) mdot (kgis] dp(psi) PR Qdot.,, I ... Y.) scroll scroll 2 fin motor pipe 1 pipe 2 pipe 3
free 1 24.82 66.92 95.27 60.07 30.50 6.18E-04 15.31 2.02 26.10 43.07 52.37 52.25 47.78 47.05 45.48 43.66 43.25
Sc-1 4 cl ,', -L *I'- E -I I 1 1 1. .-. i 4 c F l :c l, l' 1I. 1 c c I ..:I ,
r 2 1,1, -'2 I E "- I '' .- '3 4: C,, Il,, c 13 1 1 i,' | 1 '. 1 -''
,1.- I 1, .1 "' '1 E 1E I E '. : 4 1 -. cL : I -I-: c 1,, .


average 24.68 67.17 95.37 60.02 30.48 6.16E-04 15.32 2.02 26.29 43.00 51.44 52.03 46.62 47.36 44.63 44.08 42.55
std dev 0.36 0.38 0.16 0.07 0.08 1.95E-06 0.04 0.00 0.34 0.16 0.82 0.88 1.39 0.72 1.15 0.77 0.95
averages Tin Texit MC hz flow (LPM) mdot (kgis] dp(psi) PR Qdot.., q... Y) scroll1 scroll 2 fin motor pipe 1 pipe 2 pipe 3
Free 24.77 71.39 110.09 64.07 32.17 6.52E-04 16.95 2.12 30.49 42.59 53.72 53.21 48.14 49.06 46.65 46.12 44.16
ir:- 1 "1 t i ', 14 11 I ..4 i~i ) 1 I 4 I --c i F. : 1 1- r i 1 I, 14 11 .:' I
I -- 1' 1, 4 :1 -41-2 -'': 1 .E -.4 14 4i ..1 1 4 F- c- I: I : c.:' 4:. 1. -3 "
::- i .: :: i:. : ;4 C, :: ,E -'4 i ";': 4 -:::- .i i : 1i 1: 1- :. I : I .:' i il



average 24.73 72.19 109.69 64.03 32.05 6.48E-04 16.82 2.12 30.88 42.30 53.72 54.02 48.12 49.63 46.43 46.08 44.32
std de 0.35 0.61 0.29 0.03 0.08 2.91E-06 0.24 0.02 0.48 0.56 1.17 0.65 0.42 0.52 0.44 0.73 0.38
averages Tin Texit MC hz flow [LPM1 mdot (fkgsl dp(psil PR Qdot.., n... scroll 1 scroll 2 fin motor pipe 1 pipe 2 pipe 3


free 1 26.04 78.44




,,-+ ". -." -
n:- .. :-".




average 25.39 78.02
std deu 0.65 0.99
1080 rpm averages Tin Texit
Free 1 25.49 85.17
h.--- -*-' 'I


Ii-4 .._ '-


average 25.20 83.96
std dev 0.51 0.96


r


123.60


I_4l.
I... -


124.31
0.71
MC
139.14
141:1.
i. : 1


ill --
140.16
1.04


1


68.04





I: _'
68.01
0.03
hz
71.97


1- 1
" ii


72.00
0.03


33.36


-



33.38
0.04
flow (LPM)
34.86



4


34.85
0.05


6.73E-04


; ,'E-,4
--%tE --4


6.73E-04
2.15E-06
mdot (kgis)
7.04E-04



S..-E-..4


7.03E-04
1.61E-06


17.95
I. I
li. l,



r i
18.09
0.18
dp(psi)
19.19


I1



19.34
0.13


1 2.29


35.39






35.58
0.38
Qdot.,,
42.20
allf


,l '.

41.50
0.50


I


41.13


1 11%
40 :





41.14
0.44
q.,. (N)
40.19
40 -


4-,,- 4


40.10
0.34


I


59.66 59.46

'" :l f :5




58.22 58.69
1.20 1.23
scroll scroll 2
63.63 62.45



FI- :" 1 I


61.67 61.92
1.87 0.88


900 rpm











960 rpm











1020 rpm


52.86 54.58




c 414

52.34 53.57
1.20 0.99
fin motor
56.15 56.64

5-.i- f ,,




55.09 56.36
1.17 0.49


51.30
* 11


1-I'


50.18
1.44
pipe 1
53.84






52.62
1.26


52.11 1 50.17 ]


m


' '


i


-L


I


L





















300 rpm











960 rpm











1020 rpm











1080 rpm


FORCED
averages Tin Texit MC hz flow [LPM) mdot kgts]) dp[psi) PR Qdot.,, q...[.) scroll l scroll 2 fin motor pipe l pipe 2 pipe 3
forced 1 24.43 54.66 98.74 60.02 31.08 6.29E-04 15.67 2.04 19.09 43.11 39.62 37.91 34.48 39.45 26.76 27.50 24.56
:, -1. -H eq II ..E I Ic ..I I I 3 c- o 1c .
:, .- ] ._3 44 c1 3 .. In _IE ,. Ic ,, 1 ,, 3 .c 3, -,, .. t,, ._ 1 .4 1 "
i :, i-]' .c i c ,, ,,, I,, Etq ic c o I. ,, ,- I l.


IoEr I -r ..-Io 1
i:, 'I _. < l ,,I .'E Ic .,, 1 l. i i "1 3 3 Ic 3l i ._ ; .

average 25.09 55.73 98.38 60.03 31.08 6.28E-04 15.53 2.03 19.33 42.97 40.30 39.00 36.17 40.41 28.84 28.86 27.35
std dev 0.68 1.34 0.44 0.02 0.07 1.85E-06 0.12 0.01 0.47 0.39 1.63 1.18 1.67 0.83 1.39 1.45 2.48
averages Tin Texit MC hz flow (LPM) mdot (kgts] dp(psi) PR Qdot.,, q.,. ([) scroll1 scroll 2 fin motor pipe 1 pipe 2 pipe 3
forced 1 24.82 59.38 113.07 64.06 32.75 0.00 17.48 2.16 22.97 43.22 41.21 40.58 37.14 41.80 28.95 29.83 27.35
Forced 2 24.68 59.09 112.87 64.02 32.74 0.00 17.02 2.13 22.85 42.38 41.06 39.76 37.50 42.41 28.49 28.23 26.08
1: .'.. cI.. I.I -, I II .. I I .I c
1:, 1 II "II "i" 1- l ll 31 1 I 1 .1 1. 11 n I I o, 3
Forced 5 24.12 58.90 113.91 64.04 32.77 0.00 17.45 2.16 23.23 42.89 41.84 40.40 36.77 42.25 28.68 28.92 26.46
average 24.72 59.35 113.31 64.03 32.76 6.63E-04 17.35 2.15 23.07 42.90 41.16 40.05 36.92 42.38 28.34 28.39 26.26
std dev 0.36 0.39 0.46 0.02 0.02 1.65E-06 0.19 0.01 0.20 0.32 0.43 0.76 0.49 0.46 0.54 1.01 0.92
averages Tin Texit MC hz flow (LPM) mdot (kgts] dp(psi) PR Qdot.,, q.,. ([) scroll1 scroll 2 fin motor pipe 1 pipe 2 pipe 3
forced 1 25.56 63.96 128.73 68.01 34.37 6.93E-04 18.65 2.24 26.71 41.88 43.95 41.68 38.25 45.15 28.62 28.94 28.14
Forced 2 25.56 65.31 128.68 68.02 34.26 6.90E-04 18.58 2.24 27.54 41.63 44.94 42.64 39.88 45.47 32.07 30.87 29.38

i:i -. ] _." 1, 1 I 1_ I ** I 1 qE I -c -l 'l f t r 1-,,_ f- 1 l,, Ii. I.

i:, -'1 .q 1 I l .* .' I I. l .I ,, ii' l 3l ._1 "3 ll _
average 25.38 64.35 128.90 68.02 34.29 6.93E-04 18.69 2.24 27.10 41.81 44.51 42.39 39.30 45.07 30.06 30.02 28.59
std dev 0.27 0.84 0.26 0.02 0.06 1.97E-06 0.11 0.01 0.42 0.22 1.13 1.09 1.11 0.30 1.56 1.03 0.82
averages Tin Texit MC hz flow (LPM) mdot (kgts] dp(psi) PR Qdot.,, q.,. ([) scroll1 scroll 2 fin motor pipe 1 pipe 2 pipe 3
forced 1 24.97 68.72 145.48 72.07 35.97 7.27E-04 20.46 2.36 31.91 41.62 47.08 43.68 40.91 46.63 30.44 30.03 28.72
forced 2 24.45 67.70 145.24 72.06 36.01 7.28E-04 20.41 2.36 31.61 41.65 46.47 43.70 40.04 45.88 29.95 29.02 27.88
Forced 3 24.32 66.97 145.21 72.05 36.08 7.32E-04 20.66 2.37 31.34 42.16 45.50 42.54 39.99 46.69 29.37 29.52 26.24
Forced 4 24.57 67.21 145.60 72.03 36.04 7.32E-04 20.50 2.36 31.33 41.77 45.68 43.26 40.37 46.68 28.42 27.88 25.95
Forced 5 25.35 69.63 145.94 71.96 35.93 7.26E-04 20.19 2.34 32.29 41.05 46.77 45.85 40.90 46.96 31.56 32.23 30.51
average 24.73 68.05 145.49 72.03 36.01 7.29E-04 20.44 2.36 31.70 41.65 46.30 43.81 40.44 46.57 29.95 29.74 27.86
std dev 0.42 1.11 0.30 0.04 0.06 2.76E-06 0.17 0.01 0.41 0.40 0.69 1.24 0.45 0.41 1.17 1.61 1.87
























STANDARD DEVIATIONS
Sid...
STOCK
= -, )-. ,," T ,M r = i : i 1" i_ r 1 ,T 3:1lk I ,' : glk .l ]( 1 .1 I- 1 :i= "l===l"=l :11I 11 h ,' ,' : ,

,, I I ,,,- ,, ,,, I ,, 1 ,, I ,, I I I I





FREE CONVEC TION
-., i. ,, T .1, 1 : iil l : l I : l .l 1 .1 I- I : Y, 1" 1 .1:. .' 11 h ,' ,T1:. : 1.f I f 1'


"**1,T 1 li I II -1 ,t -l ,, I Il" I I I 1
h .. i 4 I l I _I--.-t .l l I I 1 I I .




FORCED CONVECTION
,. ,1 .1 1 : ii. IIF r .11 a :l I I ,': 1 :ll 1.1 l I :.1 : 1" : 11 1 :1: I i- I -
I I IE '' ''''_ I II I I
960rpm 0.86 112 114 003 0.07 2.40E-06 00087 015 0.01 072 0.40 617 138 2.12 1.51 164 1.36 2.77
I,, __ ,,' I ,,," IIE ,,,- ,, ,, ,, ,, I I I I I '
,, 7 ,'I I -I I Il ,, ,, _- I I I II II

Sid....l
STOCK
, .- ,, T,, T r =, I : 1 i F l I :1 1 1 .1 I- l :1 .,'1= 1"=1 ." :11 .W 'a 111 1i If :1



h ,_,, I, ,, ,, ,,,, l1 i- I '" '11 ,,,,_ ,, ,,, it '' 1- I'' 1 ,,,, l'1l



FREE CONVECTION
, -. ,:1 ,i, 1" 1 : .. ,,11 :.1 1V E.fO- 1 .,I'-
,,,, ',' ,, ,, I ,,,,- ,,,, I .. .... ... I ,, I ,, I .. ..I.I' 1
i ._..-r' ,1 h ,,h .. .. ,.. II IIh _,_I I_ I I II 1


I,, I ,, I,, ,,,, ,,,, I l ,, ,, ,, I ,I ,, ,, ,, I ,, I II ,



FORCED CONVECTION
,, T-Y r.. r iW IIFr. 1I 1.7 1 I .7 :1 1.1 i l .'1 1 I-:F, .,' 1" 1 :III 11 i, ,. -I f-
,' I "- I ': E I ''''I .. I II I I Il I'_



.' "i I II ,,. i .. e .,t .,.,I lei- == t i ==tI .. ,. I l t"i ll I I- I lI

























STOCK









FREE CONVECTION
,.-.,.-; T.. T .1 T r1., i.: il: iLF m i. T. v:I .,.'l.. i l| l F F F ]:0 .. i... .i m .:IIi : h,-'. ..:': I : -,_ -








FORCED CONVECTION
,.-. T.I. T- .1 T r.1. I.: ii: iLF .I11 iT ]: I .l..I ] .1 l : l : 'i... : :II : I I F -
I,,,,I,' I. .11 I 11I ,,, 1111I In ,,,, I ,, 11 "1 I, I I I I ."

I ., I S I' I l ,, l-l. 1,1,1 11 111,, "i I I I c l -I l
i r i i i ,,,,c i i,,,, ,, i
,, "I l ~ I I I. ,1, ,, II ,,,, ,, ,, -c ,, I l ,
I.-Ili 11 111 1 11 1I11 I II
FORCE COYYCTol













Uncertainty in variables due to instruments


60hz 0.010
64hz 0.040
68hz 0.020
72hz 0.010
volts 0.010
amps 0.010
amb. Press.(inHg) 0.100
dp(psi) 0.200
flow(lpm) 0.050
mdot(Kg/hr) 0.008
p(kg/mA3) 0.002
T(K) 0.050
PR 0.021
AT(K) 0.100
Qdotnet 0.118
MC 0.151
Pn=100% 0.466
lsys (%) 0.391


Final average uncertainty, co
T..I I II T- L.T I r.1 : : 11 : F IL FPr. T i lI 'I l I F I Ifi F 1 1 II.1 I :, F I
.-. ; 11 1 -. 1.-. .1 -. .1 r. ; I















APPENDIX C
MODEL SPREADSHEET SAMPLE1


1 Please refer to Chapter 5



























.i I FF. III T Ii T T T..... .. .. T. FF FF..

I I II I I ,1. 3


1 I lI,1 ._1 I.I I, 1 1 '.' I I I 111. 1. I1 I II 11 ,.'. HI





I I 1 1, ,. ~ .. ,,II 1 11 .' 1" 11 IiC I IC I' .. I ,',, 1 I I.. I ,..I .. .. I.I I "1







11 I I 1 I I 1 I I 1 I I
I I,,.. i I I-I_'. ,,,,I-. 1, I--. c- 11 ,, S. cy ,, .1.3 3




I11' 1 ~ 111 I".." I I I I II I 'I I -I 11.1 I I 1 ". 1. 11 -1 1- I. ..I- I.I.. .

I5 I,,.11.' 1 II I'.,I- I-Icc = l' I I c




I 4 I,, ,, ,r II c I.I ,, i, I I -. l c I c
Ic I Ib l ,,,,co .., II .. ". '.I 1 .el II l '. 1 I I .c IIII" 1 1 I' II I I


I 1 1 1, I. I I. .1 I -.1 ,. l q -I. I II1- -- I ,,









lIIl. I I" .5 4 ,,, i I ".. I- n"' ,,c ,,..".." I "" i '" .















APPENDIX D
MODEL GRAPHS1


1 Please refer to Chapter 5













Model Predicted Temperature Profile-Stock 900 rpm


110

100

90

80

70

60 -
50

40

30

20 ---
1 1.1 1.2 1.3 1.4 1.5 1.6 1.7
volume ratio

---Tactual -E-Tisentropic Twall




Model Predicted Temperature Profile-Stock 960 rpm


110
100
90 -
80
70
60
50
40
30
20


1.1 1.2 1.3 1.4
volume ratio


1.5 1.6 1.7


-- Tactual Tw all --- Tisentropic







79



Model Predicted Temperature Profile-Stock 1020 rpm


1.1 1.2 1.3 1.4 1.5 1.6 1.7
volume ratio


S+ Tactual --- Tisentropic


Twall


Model Predicted Temperature Profile-Stock 1080 rpm


S--Tactual --- Tisentropic


1 1.1 1.2 1.3 1.4 1.5 1.6 1.7
volume ratio








80




Model Predicted Temperature Profile-Free 900 rpm


1.1 1.2 1.3 1.4 1.5 1.6 1.7
volume ratio


-- Tactual --- Tisentropic


Twall


Model Predicted Temperature Profile-Free 960 rpm


110
100
90
80
70
60
501
40
30
20
1 1.1 1.2 1.3 1.4 1.5 1.6 1.7
volume ratio


S--Tactual -- Tisentropic


Tw all






81


Model Predicted Temperature Profile-Free 1020 rpm


1.1 1.2 1.3 1.4 1.5 1.6 1.7
volume ratio


S+ Tactual --- Tisentropic


Twall


Model Predicted Temperature Profile-Free 1080 rpm


I-7


1 1.1 1.2 1.3 1.4 1.5 1.6 1.7
volume ratio


S--Tactual -- Tisentropic


Tw all







82



Model Predicted Temperature Profile-Forced 900 rpm


1.1 1.2 1.3 1.4 1.5 1.6 1.7
volume ratio


-- Tactual --- Tisentropic


Twall


Model Predicted Temperature Profile-Forced 960 rpm


-- Tactual --- Tisentropic


1 1.1 1.2 1.3 1.4 1.5 1.6 1.7
volume ratio


Twall








83




Model Predicted Temperature Profile-Forced 1020 rpm


1.1 1.2 1.3 1.4 1.5 1.6 1.7
volume ratio


-- Tactual --- Tisentropic


Twall


Model Predicted Temperature Profile-Forced 1080 rpm


- Tactual -- Tisentropic


110

100
90
u 80
S70
60
E 50

40
30
20


110
100
90
0 80
70
60

E 50
S40
30
20


1 1.1 1.2 1.3 1.4 1.5 1.6 1.7
volume ratio


Tw all















APPENDIX E
NOMENCLATURE



7A = efficiency based on availability analysis

r7 = efficiency based on approximated shaft power

1 = efficiency based on relative compressor power

p = density

C)= uncertainty

a = model area coefficient

AA = change of availability

A, = model initial heat transfer surface area

dp = discharge pressure

Ah = change of enthalpy

H = model heat transfer coefficient

he = model calculated approximate heat transfer coefficient

hz = frequency

S= mass flow rate, mdot

MC motor controller input power

MEiosses = model predicted power loss due to mechanical forces










Motorlosses




pamb

P* ,looyo
P17 =100%


P, o =100%,m

PR

PR,

PR,

PR127


PRstage

PRstage,previous

Qnet


Qnet,m


Removed


Removed 1


Qremoved,total

rpm

AS

SP
approx

Ti

T2


= model predicted electrical and mechanical losses from

motor controller and motor

= ambient air pressure

= relative compressor power based on isentropic compression


S model predicted P=100oo

S pressure ratio

S model predicted stage beginning pressure ratio

S model predicted stage ending pressure ratio

S model final stage pressure ratio


= model predicted pressure ratio up to calculated stage

S PRtage from the proceeding stage

S net rate of heat addition, Qdotnet


= model predicted net rate of heat addition, Qdotnet,m


= model stage predicted heat removal rate, Qdotremoved

= model first stage predicted heat removal rate, Qdotremoved,1


= model predicted total heat removal rate, Qdotremoved,total

S revolutions per minute

S change of entropy

S approximated shaft power

S model predicted stage beginning temperature

S model predicted stage intermediate temperature









T3 = model predicted stage ending temperature

Tactual = model predicted compressor air temperature profile

TEx = compressor discharge air temperature

ATexit-n = difference between T,,t and T,

ATexit-wall = difference between Te,, and T,,,,

T7 = compressor intake temperature

Tisentropic = model predicted isentropic compression temperature profile

ATcro = difference between Twai, and model predicted T2

Twail = scroll wall temperature

TCP = model predicted total power used for compression

V = volumetric flow rate

VR = volume ratio

We = model predicted compressor work















LIST OF REFERENCES

1. Culp, Archie W., Principles of Energy Conversion, McGraw-Hill, Inc., New
York, 1991.

2. Incropera, Frank P., DeWitt, David P., Fundamentals of Heat and Mass Transfer,
Fourth Edition, John Wiley & Sons, Inc., New York, 1996.

3. Larminie, James, Dicks, Andrew, Fuel Cell Systems Explained, Second Edition,
John Wiley & Sons, Inc., New York, 2003.

4. LG, Technical Manualfor LG Scroll Compressor, Version 1, LG Electronics Inc.,
Air Conditioning Compressor Division, Englewood Cliffs, NJ.

5. Michael J., Shapiro, Howard N., Fundamentals of Engineering Thermodynamics,
Third Edition, John Wiley & Sons, Inc., New York, 1996.

6. Peterson, G.P., An Introduction to Heat Pipes, Modeling, Testing, and
Applications, Wiley-Interscience, New York, 1994.

7. Radermacher, R., Schein, C., "Scroll Compressor Simulation Model," Journal of
Engineering for Gas Turbines and Power, Vol. 123, January 2001, p217-225.

8. Silverstein, Calvin C., Design and Technology of Heat Pipes for Cooling and
Heat Exchange, Hemisphere Publishing Corporation, Bristol, PA, 1992.