ins?!»
529:1 m nfl
at...

.s
~ .22
1:3 .5.
.iictuuflu.
.33.... 1‘3; :5
B... ) in: I
I! tel {I .K (1223*
dcrflufi .53....“ r
1e ‘ c
2:55-12:13

nave-v.1;

..

3:.
any». .sfigw
.. .nn...::.is

fl "$9.251:

, I: 3
nbmgemfi. brad.»
. .A :; IJMIMNUAHH 0H?I3Q
.35! 3...: a. t. .
.3" . :53: a! s.
. LL! 5%,???
.
. . a
3 1...:
13;...
z 1
17.0%....

3:.
51-01}: It
(5.3.5, I.)
.29.th
3, 3.1.1:!!!
32.11:..ldllahuirr
oi . rififil...

z)
w. x1: . .
5:... cw! . ‘555
xt:......rcI’/.
3 l. .0 no.3:
. . n11:
.4

. ‘ I L7: .5.
l:- .|ll.s|. , 4.33:»!
V j . r 3.22.. .
nil-.3. \ .6 :23? z .
3.1.; 4J5 ; zé :fiwwumfi
Jill: ‘. f ::I 11.1., . .5.

1...)“.1 flaunt. .- ”I”... . :7 1;.
.<oll'.1"\l’\.. . I. . ......v
33.3.
.3

 

This is to certify that the
thesis entitled

STUDY OF A NOVEL R718 TURBOCOMPRESSION CYCLE

presented by

Amir Ahmadzadeh Kharazi

has been accepted towards fulfillment
of the requirements for the

Ph.D. degree in Mechanical Engineerirfl

 

 

(1/ £6

 

Major Professor's Signature
06725106

Date

 

MSU is an Afiinnativo Action/Equal Opportunity Institution

,,f ,, ——— w»-

 

LIBRARY
Michigan State
University

 

 

 

.-.-.----n--a--n-n-n---n-0---sQI-o-o--o-c-‘-u-o-0---o-o-o---~-c.

.-.-o-n-o-o-o---o-o-o-n-.-.-.-.- —.-

 

PLACE IN RETURN BOX to remove this checkout from your record.
TO AVOID FINES return on or before date due.
MAY BE RECALLED with earlier due date if requested.

 

DATE DUE DATE DUE DATE DUE

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

2/05 p:lClRC/DateDue.indd-p.1

STUDY OF A NOVEL R718 TURBOCOMPRESSION CYCLE
By

Amir Ahmadzadeh Kharazi

A DISSERTATION
Submitted to
Michigan State University
in partial fulfillment of the requirements
for the degree of
DOCTOR OF PHILOSOPHY

Department of Mechanical Engineering

2006

ABSTRACT

STUDY OF A NOVEL R718 TURBOCOMPRESSION CYCLE USING A
3-PORT CONDENSING WAVE ROTOR

by
Amir A. Kharazi

Water is a natural refrigerant, which is absolutely harmless to man and
nature. It is easily available and incurs no problems with its disposal after use.
Even though water is one of the oldest refrigerants, state of the art technology is
required to use water as a refrigerant in compression refrigeration plants with
turbocompressors.

To compare water (R718) to other refrigerants, a code is developed in which
all refrigerants can be compared in a single p-h, T-s, or p-T diagram. Using the
code, the COP isolines of water (R718) and any refrigerant can be generated in a
graph to determine which refrigerant has a better COP for a certain evaporation
temperature and temperature lift. In regard to using water (R718) as a refrigerant,
some specific features complicate its application in refrigeration plants with
turbocompressors. Because the cycle works under coarse vacuum, the
volumetric cooling capacity of water vapor is very low. Hence, huge volume flows
have to be compressed with relatively high pressure ratios. Therefore, the use of
water (R718) as a refrigerant compared to classical refrigerants, such as R134a
or R12, requires approximately 200 times the volume flow, and about twice the

pressure ratio for the same applications.

To enhance the turbocompression and improve the efficiency of R718 cycles,
the novel concept of 3-port condensing wave rotors integrated in R718
compression refrigeration cycles is investigated. The condensing wave rotor
employs pressurized water to pressurize, desuperheat, and condense the
refrigerant vapor, all in one dynamic process. The underlying phenomena of flash
evaporation, shock wave compression, desuperheating, and condensation inside
the wave rotor channels are described in a wave and phase-change diagram as
part of internal flow analysis. A characteristic equation is developed for the
determination of the pressure required for the high pressure inlet port when the
low pressure port and the outlet are known. For the external analysis, the
thermodynamic process is described in detail. Based on the described
thermodynamic model, a computer program is generated to evaluate the
performance of the R718 baseline and wave-rotor-enhanced cycles. The effect of
some key parameters on the performance enhancement is demonstrated as an
aid for optimization. A performance map summarizes the findings and shows
optimum wave rotor pressure ratio and maximum relative performance
improvement of R718 cycles by using the 3-port condensing wave rotor.

Experimental investigation of the flow inside a small size wave rotor is
described in the last chapter. This experimental analysis is reported as a proof of
concept for wave energy exchange in smaller sizes, which ultimately lead to
micro wave rotors. Moreover, the Schlieren method is used for shock wave
visualization. Based on the skills developed, Shlieren technique is

recommended for flow visualization in channels of a condensing wave rotor.

TO MY WIFE

ACKNOWLEDGMENTS

I am taking this opportunity to thank all those who have assisted me in one
way or the other with my PhD. thesis.

I would like to thank professor Norbert Mueller for giving me the opportunity to
work with the Turbomachinary group at the Michigan State University. The
friendly and supportive atmosphere inherent to the whole Turbomachinary
labratory contributed essentially to the final outcome of my studies. In this
context, I would like to thank particularly for my Ph.D. committee professors
Jaberi, Engeda, and MacCluer. Also, I would like to thank Pejman Akbari for his
guidence during early stages of my Ph.D., with whom I published several
publications.

For their hospitality, I would like to thank the wave rotor group at the University
of Tokyo, where the experimental study of this thesis was conducted.
I am particularly grateful to professor Charles Petty, who was so kind as to give
me opportunities in the Center for Multiphase Transport Phenoma. Without his
support the completion of this dissertation would not have been possible.

Expressing gratitude to persons who played a role during a PhD. thesis can
be an unending task for even minor or seemingly unrelated contributions may
affect or influence it substantially. Therefore, I would like to thank my parents,
whose contn'bution was not related directly to the work itself but without knowing
of their moral support and assurance, their found conviction in principle question

of life and their strong belief, this work would not have been possible. I truly owe

everything to them, my gratitude and acknowledging their choices and decisions,
which build the basis for a unique family.

Finally, I owe special thanks to my wife Mahnaz for her patience, love, and
encouragement during these years of my studies.

This work received financial support from the National Science Foundation,

MSU Graduate Office, and MSU Engineering Department.

VI

TABLE OF CONTENT

LIST OF TABLES ............................................................................................... IX
LIST OF FIGURES .............................................................................................. X
CHAPTER 1: INTRODUCTION ............................................................................ 1
1.1. REFRIGERATION ................................................................................................ 1
1.2. NATURAL REFRIGERANTS ................................................................................ 3
1.3. WATER AS A REFRIGERANT ............................................................................. 8

CHAPTER 2: COMPARATIVE INVESTIGATION OF WATER AS A

REFRIGERANT .................................................................................................. 17
2.1. COMPARISON IN A TEMPERATURE PRESSURE DIAGRAM .......................... 17
2.2. COMPARISON IN A TEMPERATURE ENTROPY DIAGRAM ............................ 19

2.2.1 DESUPERHEATING AND THROTTLING IRREVERSIBILITY FACTORS ................. 20
2.3. COMPARISON IN A PRESSURE ENTHALPY DIAGRAM .................................. 25
2.4. COEFFICIENT OF PERFORMANCE COMPARISON ........................................ 28

2.4.1 COMPARING COP OF WATER (R718) AND R134a ................................................. 28

2.4.2. COMPARING COP OF WATER (R718) AND AMMONIA (R717) .............................. 30

CHAPTER 3: IMPLEMENTATION OF A CONDENSING WAVE ROTOR ......... 32
3.1. WAVE DEVICES ................................................................................................ 32
3.2. WAVE ROTOR ................................................................................................... 34
3.3. APPLICATION OF WAVE DEVICES IN REFRIGERATION CYCLES ................ 39
3.4. CONDENSING WAVE ROTOR .......................................................................... 40

CHAPTER 4: ANALYSIS OF CONDENSING WAVE ROTORS ........................ 44
4.1. INTERNAL: TRANSPORT AND PHASE CHANGE PHENOMENA ..................... 44

4.1.1 STUDY OF FLOW INSIDE A CHANNEL OF A CONDENSING WAVE ROTOR ........ 48
4.1.1.1. TEMPERATURE INCREASE ACROSS THE NORMAL SHOCK WAVE ........... 48
4.1.1.2. INTERFACE VELOCITY ACROSS THE NORMAL SHOCK WAVE ................... 50
4.1.1.3. CHARACTERISTIC EQUATION ......................................................................... 52

4.2. EXTERNAL: THERMODYNAMIC CYCLE .......................................................... 54

VII

4.3. SIZE COMPARISON .......................................................................................... 58

5. PERFORMANCE EVALUATION OF CONDENSING WAVE ROTOR

INTEGRATED IN A R718 CYCLE ...................................................................... 60
5.1. THE BASELINE PERFORMANCE CALCULATION ............................................ 61
5.2. THE WAVE ROTOR ENHANCED CYCLE PERFORMANCE CALCULATION ...63
5.3. PERFORMANCE RESULTS AND DISCUSSIONS ............................................. 68

5. EXPERIMENTAL STUDY OF SHOCK WAVE VISUALIZATION USING

SCHLIEREN METHOD ....................................................................................... 75
5.1. SCHLIEREN VISUALIZATION TECHNIQUE ...................................................... 75
5.2. SCHLIEREN TECHNIQUE FOR VISUALIZATION OF SHOCK WAVE IN A
SMALL SIZE WAVE ROTOR .................................................................................... 76

SUMMARY ......................................................................................................... 82

REFERENCES ................................................................................................... 84

VIII

LIST OF TABLES

Table 1: Required compressor pressure ratio for several refrigerants ................ 18

Table 2: Comparison of irreversibility and throttling factors for several refrigerants

for Tevap._=5°C and Tcond._=35°C ......................................................... 24
Table 3: Comparison of different refrigerants in a typical evaporation and
condensation temperature .................................................................. 27
Table 4: Marginal conditions for the performance comparison in Figure 14, taken
from Ref. [27] ...................................................................................... 29
Table 5: Input data for the baseline cycle analysis .............................................. 61
Table 6: Input data for the enhanced cycle analysis ........................................... 63

Table 7: Calculated baseline cycle state values for the selected optimum point P
............................................................................................................ 72

Table 8: Calculated enhanced-cycle state values for the selected optimum point P
............................................................................................................ 72

LIST OF FIGURES

Figure 1: Basic compression refrigeration cycle ................................................... 2
Figure 2: The German Deutche Bahn ICE-1 train, air conditioned with
compressed air as refrigerant ............................................................... 4
Figure 3: Schematic of a R718 chiller unit with direct condensation and
evaporation ......................................................................................... 1 1
Figure 4: Schematic of thermodynamic model of a R718 chiller unit with a Mo-
stage compressor and a direct condenser .......................................... 11
Figure 5: 100-300 ton chillers commercialized in Europe ................................... 15
Figure 6: Pressure-Temperature diagram for common refrigerants .................... 17
Figure 7: Pressure-Temperature diagram for common refrigerants for a common
refrigeration operating range ............................................................... 18
Figure 8: Temperature-Entropy diagram for common refrigerants ...................... 19

Figure 9: Entropy production dun'ng desuperheating and throttling for R718 and R134a 20

Figure 10: Slope of a constant pressure line in Temperature-Entropy diagram ..23

Figure 11: Throttling irreversibility factor in a T-s diagram .................................. 24
Figure 12: Log pressure-Enthalpy diagram for R718, R717, and R134a ............ 26
Figure 13: Temperature-Enthalpy per specific volume diagram .......................... 27
Figure 14: COP isolines for R718 and R134a refrigerants .................................. 28
Figure 15: COP isolines for R718 and R717 refrigerants .................................... 30
Figure 16: Schematic configuration of a typical wave machine [43] .................... 33
Figure 17: Comparison of pressure gain (static pressure ratio) of moving shock

and steady flow isentropic diffuser for y=1.4 [42] ................................ 34
Figure 18: Shock wave and diffuser isentropic efficiencies as a function of the

static pressure ratio [42] ...................................................................... 35
Figure 19: Schematic configuration of a typical wave machine ........................... 37

Figure 20: A schematic of a gas turbine topped by a 4-port wave rotor [42] ....... 38

Figure 21: Dynamic-pressure equalizer .............................................................. 41
Figure 22: Schematic of a 3-port condensing wave rotor .................................... 43

Figure 23: Schematic of the thermodynamic model of a R718 chiller unit
enhanced by a 3—port condensing wave rotor (CWR) substituting for
the condenser and for one stage of compressor ................................. 44

Figure 24: a) Schematic wave and phase-change diagram for the 3-port
condensing wave rotor (high-pressure part) b) A magnified channel
showing the regions modeled during compression and condensation.

............................................................................................................ 45
Figure 25: Schematic wave and phase-change diagram for the 3-port condensing
wave rotor including the discharge process ........................................ 47
Figure 26: Regions modeled during compression and condensation as described
in section 4.1 ....................................................................................... 48
Figure 27: Flash evaporation photography by Miyatake [65] .............................. 52
Figure 28: Condensing wave rotor wave diagram ............................................... 53

Figure 29: Schematic p-h diagram of a R718 baseline cycle and enhanced cycle
with a 3-port condensing wave rotor. .................................................. 55

Figure 30: Schematic T-s diagram of a R718 baseline cycle and enhanced cycle
with a 3-port condensing wave rotor ................................................... 56

Figure 31: Novel compact R718 water chiller with integration of a condensing
wave rotor ........................................................................................... 57

Figure 32: Size reduction by combining intercooler, second stage compression,
and condensation into a condensing wave rotor (CWR) ..................... 58

Figure 33: Relative COP increase versus evaporation temperature for different
mass flow ratios .................................................................................. 68

Figure 34: Relative COP increase versus mass flow ratio for different evaporation
temperatures ....................................................................................... 69

Figure 35:Relative COP increase versus the wave rotor pressure ratio for
different mass flow ratios. ................................................................... 70

Figure 36: Heat rejecter temperature versus evaporator temperature for different
wave rotor pressure ratios. ................................................................. 71

Figure 37: Performance map: maximum performance increase and optimum
wave rotor pressure ratios. ................................................................. 74

XI

Figure 38: Diagram of Schlieren setup for shock visualization ............................ 76

Figure 39: Schlieren setup at University of Tokyo for shock wave visualization in

a single channel of a wave rotor ......................................................... 77
Figure 40: Especial design for fluid study in a single channel of a wave rotor
(University of Tokyo) ........................................................................... 78
Figure 41: A 38 mm Glass-made channel for Shock Visualization ...................... 78

Figure 42: Pressure distribution at the end wall of the short and long test sections
(rr =25, A= 3*3) .................................................................................. 79

Figure 43: Shock wave visualization using Schlieren method for a 38 mm test
section (1r =2.5, N=40 Hz, A= 3*3) ...................................................... 80

Figure 44: Shock wave visualization using Schlieren Method for 165 mm Test
Section ((11 =2.5, N=40 Hz, A= 3*3) .................................................... 81

XII

CHAPTER 1: INTRODUCTION

1.1. REFRIGERATION

Refrigeration and Air conditioning encompasses every sphere of human life. It
not only provides comfortable and healthy living environments, but has also
come to be regarded as necessities for surviving severe weather and preserving
food. Unfortunately, accelerated technical development and economic growth in
much of the world during the last century have produced severe environmental
problems so that in the mid nineteen-eighties the world recognized that CFC
refrigerants, once considered safe for people as well as the planet, were in fact
severely damaging the Earth’s ecology. Refrigerants suddenly went from being
rarely discussed to a hot topic of research. Aside from cost reduction, being
environmentally safe is being the biggest driving forces for technical innovation
and using natural refrigerants for refrigeration.

Refrigeration is defined as “the transfer of heat from a lower temperature
region to a higher temperature one” [1]. Refrigeration devices that produce
refrigeration operate using the vapor-compression cycle. Heat pumps,
refrigerators, automotive air-conditioners, and residential or commercial air-
conditioners are examples of refrigeration devices. All of these devices have one
thing in common, to reduce the temperature of an enclosed environment.

The purpose of the refrigerator is to remove heat from the cold region while
requiring as little external work as possible. A measure of the efficiency of the
device is therefore coefficient of performance (COP) which is equal to the rate of

heat removal from the cold region to the rate work is done:

cop=l (1)

w
The most common refrigeration cycle in use today is the vapor refrigeration
cycle. As shown in Figure 1, the vapor compression cycle uses energy input to
drive a compressor that increases the pressure of the refrigerant which is in the
vapor state. The refrigerant is then exposed to the hot region in a device called a
condenser. As a result, heat is transferred from the refrigerant to the hot region
causing it to condense. The refrigerant then passes through the expansion valve
across which its pressure and temperature drop considerably. The refrigerant
temperature is now below that which exists in the cold or refrigerated region in
the device called an evaporator. As a result, heat is transferred from the
refrigerated region to the refrigerant causing it to pass from the liquid state to the
vapor state again. The refrigerant then again passes to the compressor in which

its pressure is increased and the whole cycle is repeated.

   

‘ ‘
r 7

expansion valve
Figure 1: Basic compression refrigeration cycle

1 .2. NATURAL REFRIGERANTS

In 1974, scientists published a laboratory study on ozone layer depletion by
chlorofluorocarbon (CFC) refrigerants, solvents, aerosol product propellants and
foam blowing agents [2]. Further studies estimated that the ozone layer would be
depleted by CFCs by about 7% within 60 years. Based on such studies the US
has banned CFCs in aerosol sprays since 1978. By the mid-1980s, the overall
concern about CFCs was raised again when scientists discovered that CFCs
were also efficient greenhouse gases. In 1984-1985, Japanese and British
scientists discovered that the ozone layer above the Antarctic was dangerously
depleted. Therefore, urgent action was needed to prevent increasing skin cancer,
cataracts, suppression of the human immune system, and destruction of crops
and natural ecosystems [3]. United Nations Environment Program (UNEP)
coordinated the unprecedented response, involving scientists, engineers,
industry, environmental pressure groups and the public, as well as international
agencies, governments and diplomats. In 1985, the first general ozone
agreement, known as the Vienna Convention for the Protection of the Ozone
Layer, was a pledge to protect the ozone layer [4]. The specific commitments
came in 1987 through the Montreal Protocol signed by fifty seven industrial
nations. The 1987 Montreal Protocol has been progressively strengthened
several times for the rapid phase-out of the production and consumption of ozone
depleting substances [5].

Since the Montreal Protocol, there has been a tremendous effort among the

refrigeration and air-conditioning industry to find the best substitute for CFCs now

under phase out [6,7]. The search for new and environmentally benign
refrigerants has renewed interest in natural refrigerant technologies. Although
natural refrigerants were used extensively in the early years of refrigeration
technology, a number of technical and safety challenges caused them to be
readily abandoned when CFCs became available. These challenges still exist
today for air. ammonia, hydrocarbons, carbon dioxide, and water.

Air (R729) is used extensively as a refrigerant in air-conditioning in aircrafts
and train compartments (Figure 2). The use of air as a refrigerant is based on the
principle that when a gas expands isentropically from a given temperature, its
final temperature at the new pressure is much lower. The resulting cold gas, in
this case air, can then be used as a refrigerant [8]. Air-cycle systems are
compact and lightweight. However, the efficiency of such systems is quite poor
and is mainly limited by the efficiencies of compression and expansion devices,
as well as those of the heat exchangers employed. The air refrigeration cycles

appear to use more energy than comparable vapor compression cycles [9].

  

”'«4.

Figure 2: The German Deutche Bahn ICE-1 train, air conditioned with
compressed air as refrigerant

Ammonia (R717) has been the most widely used natural refrigerant. After
more than a century using ammonia as a refrigerant, a tremendous amount of
practical experience exists with this refrigerant. There is no doubt about its
superior thermodynamic and transfer properties compare to other available
refrigerants. However, because of its high toxicity and flammability, thorough
care should be paid in handling and such systems require more preventive
maintenance [10]. For instance, the US Occupational Health and Safety
Administration (OSHA) considers ammonia as a hazardous material, and
imposes certain regulations on its use, storage, handling and occupational
exposure [11]. As a result, the applications of ammonia are mostly limited to
large warehouses where it can be strictly managed such as at a plant.

Hydrocarbons (HC) are appropriate refrigerants in many ways including
energy efficiency, critical point, solubility, transport, and heat transfer properties
[12]. For refrigeration applications, a number of hydrocarbons such as propane
(R-290), n-butane (R-600) and isobutene (R6003) are suitable to use. However,
they are also flammable and explosive, which causes the need for changes in
standards and care in their production. The use of hydrocarbons in North
America is very limited; because of restrict laws for flammability risks. However,
some countries in Europe and elsewhere have less-stringent liability laws.
Nevertheless, the danger of fire remains an overriding concern. To address this
challenge with safety features, the cost of a system would have to be increased

by about one-third [13].

Carbon Dioxide (R744) has very low toxicity and operates at very high
pressures which sound promising for systems that must be small and light-
weight, such as automotive or portable air conditioners. Carbon Dioxide systems
are “transcritical” meaning the compressed gas out of compressor must be
cooled against a heat sink temperature that is usually above the critical
temperature [14]. Because of the nature of the transcritical cycle, the efficiency of
Carbon Dioxide is quite poor (low specific cooling capacity) requiring more
energy for compression [15].

Water is a natural and clean refrigerant with no Ozone Depletion Potential
(ODP=0), no contribution to the global warming (GWP=O), and without any safety
risk. It is chemically stable, nontoxic, nonflammable, commonly available, and
easily disposable after use. In R718 systems, water serves as both refrigerant
and heat transfer fluid. Therefore, direct heat exchangers can be used for
evaporator and condenser increasing the energy-saving potential. Moreover,
water can function as a working fluid for refrigeration and heat pump cycles in the
range of evaporation temperature of 0 °C to very high temperatures (e.g. 150 °C).
Even though water is one of the oldest refrigerants, state of the art technology is
required to use water as a refrigerant in absorption chillers or in compression
chillers with steam injection compressors. These techniques are energy efficient
as long as there is waste heat or steam available. Much higher energy efficiency
can be achieved using compression refrigeration plants with turbocompressors.

Water has been used as a refrigerant for air-conditioning and for the

production of ice, using steam ejectors, large centrifugal compressors, and large

high-speed axial compressors. The future of water as a refrigerant for vapor
compression refrigeration depends either on the innovative designs of the cycle
or development of high-speed axial compressors to operate at the very low
pressures required for R718 refrigerating.

The key component of a R718 turbochiller is the compressor, since water as
a refrigerant has some specific features that complicate its application in
refrigeration plants with turbocompressors. Since the cycle works under coarse
vacuum, the volumetric cooling capacity of water vapor is very low. Hence, huge
volume flows have to be compressed with relative high pressure ratios.
Therefore, the use of water (R718) as a refrigerant ,compared to classical
refrigerants, such as R134a or R12, requires approximately 200 times the
volume flow, and about twice the pressure ratio for the same applications.
Because of the thermodynamic properties of water vapor, this high pressure ratio
requires approximately a two to four times higher compressor tip speed,
depending on the impeller design; while the speed of sound is approximately 2.5
times higher. Reynolds numbers are about 300 times lower and the specific work
transmission per unit of mass has to be around 15 times higher.

This states challenges for the compressor design. Today they are
successfully solved in commercial industrial plants mainly installed in Europe
using unique high-performance mixed-flow turbocompressors with or without
stationary guide vanes. Other concepts are under investigation, like mixed flow
compressors with inducer or pre-runner or axial multistage compressors,

promising a higher pressure ratio or a more compact design. High pressure ratios

are obtained by the combination of high rpm and large diameter, where the
diameter is primarily limited by the available space and manufacturing facility.
Working under vacuum the forces exerted by the gas on blades are very small.
So the blades must primarily withstand centrifugal forces resulting from their
inherent mass allowing for more economic lightweight constructions with
extremely thin, mostly straight blades, made of special materials like titanium or
composites. These impellers are very much different from usual high-
perfonnance impellers. They cannot be milled. They are built of several parts,
indicating a challenge for manufacturing and balancing of the high-speed system.
Advantageous is the use of variable speed direct-drive motors, working under
vacuum and water vapor atmosphere, requiring special bearing, cooling and
electrical isolation. Such motors are not readily available and need to be
developed for each new power level. Furthermore R718 turbochillers have a
lower noise emission and require no special safety installation concerning
drainage and ventilation. R718 turbochillers represent a cutting edge technology,
which is challenging and rewarding. They allow a move towards greener
technology with a huge potential for application and manufacturing even outside
of Europe [16]. If R718 cycles are successfully developed, one can imagine that
many toxic and polluted sources, such as urban waste water, could be a valuable

source of energy for refrigeration or heat pumps.

1.3. WATER AS A REFRIGERANT
There has been a tremendous effort by the refrigeration and air-conditioning

industry to find the best substitute for chlorofluorocarbon (CFC) refrigerants [6,

7]. The search for new and environmentally benign refrigerants has renewed

interest in technologies that use natural refrigerants, such as water (R718).

Considering all pros and cons of natural refrigerants already described in

previous studies [17, 18], water can be considered as an attractive refrigerant

because of its following advantages:

It has no global warming potential (GWP=O).

It has no ozone depletion potential (ODP=O).

It is nontoxic, nonflammable, easy to handle, and inert to the environment
(minimizes safety precautions).

It has no risk of future restrictions due to refrigerant environmental impact.

It creates no disposal problem after use.

It works with very low pressure differences, reducing safety precautions.

It has high theoretical coefficient of performance (COP), competitive with
CFCs depending on the evaporation temperature [19, 20].

The system working with water as a refrigerant can use direct heat
exchangers for evaporation and condensation. Therefore, R718 systems can
obtain very high COP [21].

Tap water, treated waste water, or coarsely filtered river water can be used
directly as make-up water (warehousing bulky refrigeration canisters is not
required).

Chiller systems, coupled with a closed cooling-tower loop, allow for diminished

water treatment.

- Turbochillers using water as a refrigerant have shown to be inherently much
less noisy than conventional compression chiller systems.

Despite the above attractive features, there are a few challenges of using
water as a refrigerant compared to traditional refrigerants. At the triple point, the
vapor pressure of water is only 611 Pa, which < 1% of the atmospheric pressure.
The low operating pressures of water-vapor refrigeration systems combined with
the steep vapor pressure curve of water requires compression systems that can
handle large-volume flows while still delivering high-pressure ratios [22-25]. This
states challenges for the compressor design. Although single-stage
turbocompressors commonly deliver large-volume flows with mostly insufficient
pressure ratios, positive displacement compressors can obtain high-pressure
ratios but only for relative small-volume flows. A technical compromise has been
the use of multistage turbocompressors with intercooler [26]. The drawback of
multistage compressors with intercoolers is being spacious.

For the reasons described, R718 hardly competes with other commonly used
refrigerants in small household size refrigeration or air-conditioning applications.
Any affordable cycle modification or newly invented device which could
simultaneously handle large volume flows while still achieving high pressure

ratios will give R718 an edge over other commonly used refrigerants.

10

 

 

 

 

 

Figure 3: Schematic of a R718 chiller unit with direct condensation and
evaporation

Figure 3 pictures the schematic of a two-stage R718 turbochiller with direct
condensation and evaporation, two high-perfomiance centrifugal compressors,

and an intercooler. Figure 4 is a schematic thermodynamic model for the actual

system.

   

throttling valve

 

 

  

 

calms man

9|vo mm pailiuc

  

Cooling water cycle

  

 

 

w“ throttling valve

Figure 4: Schematic of thermodynamic model of a R718 chiller unit with a two-
stage compressor and a direct condenser

11

R718 units can be comprised of three directly interlinked cycles: (i) a cooling
water cycle that condenses the superheated vapor from the compressor and
releases the thermal energy to the ambient through a heat exchanger device,
mostly a cooling-tower; (ii) a chilled water cycle that absorbs thermal energy from
the heat source and transfers it to the refrigerant by phase change; (iii) a
refrigerant cycle or the core cycle that consists of four components: compressor,
expansion device, condenser, and evaporator. Cooling-towers can be direct
(open-circuit) or indirect (closed-circuit) heat rejection equipment. In the direct
type, cooling water is exposed directly to the atmosphere. The warm cooling
water is sprayed over a fill that increases the contact area, and air is blown
through the fill. The majority of heat removed from the cooling water is because
of partial evaporation of the cooling water requiring a permanent replenishment.
The remaining cooled water drops into a collection basin and is recirculated to
the chiller. Using such an open circuit, a pump is required before the cooling-
tower to pump the water out of condenser at subambient pressure to somewhat
above ambient pressure. In this case, considerable throttling occurs after the
cooling-tower when the water enters the condenser under vacuum, which is
shown by a throttling valve in Figure 3. The use of direct heat exchangers may
result in contamination of the refrigerant with noncondensable gases, solids, or
entrained liquids [24]; therefore, using indirect heat exchangers may reduce the
cost associated with degasifiers and cleaning. There are two options for
realization, which can be used alternatively or combined. First, an indirect

cooling-tower can be employed in which water circulates through tubes (coils)

12

located in the tower without coming in contact with the outside air and pressure.
The cooling-tower heat transfer still may be enhanced as needed by wetting the
outside of the coils utilizing evaporative cooling. If an indirect cooling-tower
(closed circuit) is used, the pump needs only to overcome the pressure loss in
the cooling circuit and throttling losses at the entrance to the evaporator can be
minimized. In such a case the pump can be placed before or after the cooling-
tower, depending on where pump cavitation is prevented best considering the
competing effects of temperature drop and pressure loss across the cooling-
tower. Second, the cooling-tower can be separated from the chiller unit using
effective plate heat exchangers. These typically introduce only 1 K additional
temperature difference that adds to the ~1 K temperature difference in the
internal direct heat exchanger (condenser). This is still superior to an internal
indirect condensing heat exchanger, which may typically introduce a temperature
difference of ~5 K.

Disadvantages of the current state-of-the-art R718 units (Figure 3) are mainly
their size and cost. The size is because of the use of two centrifugal compressors
with comparably large diameters and voluminous (same diameter) internal direct
heat exchangers. The cost is mainly generated by the two relative expensive
compressors with independent variable speed drives. An additional challenge
has been obtaining high peak compressor pressure ratios for high temperature-
lifts. One reason has been the limited compressor efficiency that results from the
special wheel construction and the limited pressure recovery of the steady-state

diffuser which decelerates the high-speed vapor flow out of the high-performance

13

wheel. Low Reynolds numbers of water vapor under vacuum (300 times lower
than if R1343 or R12 are used) [25] are an additional challenge for achieving a
high efficiency.

Although water as a refrigerant has been known for a long time, related
research and commercial use has been exploited only recently, using mechanical
vapor compression mechanism [28]. Compressors [29], considered for water
vapor compression technology, are usually axial compressors [30], centrifugal
compressors, cycloid compressors

[31], roots and liquid ring compressors [32], and jet ejectors [33,34]. As an
example for the use of axial compressors the HydroFrioT", chiller has been
introduced by INTEGRAL [35]. HydroFrioT" was projected to consist of a flash
evaporator, direct contact condenser, and an axial compressor, which is used as
water chiller, ice water chiller, binary ice generator (VacuumlCE©), heat pump,
and cascades. IDE (IDE Technologies Ltd.) of Israel has been the leader in
commercial centrifugal water vapor compression technology. IDE often uses two
stage compressors to reach the required temperature lift and pressure ratio [36].
With financial support from the Danish Energy Agency (DEA), the Sabroe
Refrigeration Company and DTI have used the IDE centrifugal compressors to
build a refrigeration plant, using solely water as refrigerant in the compression
process [37]. INTEGRAL has also used IDE centrifugal compressors for research
and installations [35]. The IDE centrifugal compressors have been developed into
two products called ECOVIM and ECO-CHILLER. The ECO-VIM (Vacuum Ice

Machine) is used to condense the water vapor to ice [38]. The ECO-CHILLER

14

 

removes heat from the water by direct flashing to producing chilling [39]. Figure 5

shows examples of chillers with 100-300 ton capacity commercialized in Europe.

  
  

 
   

1 X R718—TurbO-Chiller 4 x R718-Turb0—Chiller DaimlerChrysler,
GAE. Guardian Luxemburg (2001) Facility Dusseldorf (2001 , 2003)

 

I x R718Turbo43hiller. ' ' ' 2 x R718-Turbo—Chiller .Transparenr Factory:
Essen University {2000) Volkswagen Dresden (2001)

 

 

Figure 5: 100-300 ton chillers commercialized in Europe

Following an intensive multi—year developing effort, lnstitut filr Luft— und
Kaltetechnik (lLK) Dresden, over the last years has successfully commercialized
its modern chillers with now multiple units operating in Europe such as those at
the Volkswagen and DaimlerChrysler factories in Germany and GAE Gardian
Luxemburg. To the best knowledge of the author, all commercial
turbocompression systems using water as a refrigerant built so far have been in

large scale and a need for smaller units, in an economical and efficient solution,

15

 

exists. The presented work is part of the ongoing research to improve the
compressors and heat exchangers and to reduce the size and cost of the whole

unit.

16

CHAPTER 2: COMPARATIVE INVESTIGATION OF WATER AS A
REFRIGERANT

2.1. COMPARISON IN A TEMPERATURE PRESSURE DIAGRAM

Figure 6 compares water to other common refrigerants in the pressure-
temperature diagram. Water has a very high critical point compare to other
refrigerants, which makes it a competitive refrigerant in heat pump applications.
One restriction when using water as a refrigerant is that it is restricted to
applications in which the cooling requirements are above its freezing point (0°C),

if no antifreeze additives are added.

 

     

25000 .
p-T diagram ,
20000 i ' ' ' .
E
a 15000 .. H
n' : . g R71-r
5000 ’ R295? R123 .

 

 

 

o " ' i ' ' 'I ‘ "" " '
-100 -50 0 50 100 150 200 250 300 350 400
1'rc1

Figure 6: Pressure-Temperature diagram for common refrigerants

Figure 7 is again the Pressure-Temperature diagram zoomed in common
refrigeration operating range. As shown in this graph, water works in sub-ambient

pressure and its p-T curves lies under all other investigated refrigerants.

17

Although water works with very low pressure differences, its low pressure
operating point cause higher pressure ratios in the compressor. A rule of thumb
for refrigerants is that as lower the evaporator pressure, as higher the pressure
ratio of the compressor.

Table 1 shows the required compressor pressure ratios for several
refrigerants in a simple compression refrigeration cycle. R718 requires 2 to 3
times higher compressor pressure ratio when is used in a typical compression

refrigeration cycle.

 

 

 

 

 

 

0 5 10 20 25 3O 35

15
T [°CI

Figure 7: Pressure-Temperature diagram for common refrigerants for a common
refrigeration operating range

 

- v WW 0 ..‘w
-' - 'l. .‘-A,‘.~r.‘l: . x; ”.‘or-.‘.’

.. 0'1“ ‘ '
v I I I I
. . 0 -
i e erant '
A. . , I ‘ , p , _ . .
. . . j ‘ - II“ P,- - < "'23 ‘ .."T '. Zl't

. ,.,..... , ,
. 223133.“): 6: p14,, [7

  

 

 

{v CempressiOn Ratio ‘ 5.

:(T= 5°C,T= 35°C) ‘ 2.3 2.2

 

 

 

 

 

 

 

Table 1: Required compressor pressure ratio for several refrigerants

18

2.2. COMPARISON IN A TEMPERATURE ENTROPY DIAGRAM

Figure 8 is the comparison between investigated refrigerants in a
temperature-entropy diagram. As shown in the picture, water (R718) has a wide
dome compare to the other refrigerants. Therefore, when used in a refrigeration
cycle during evaporation and condensation, entropy change is significant. This
means, for example in the evaporator, water molecules need a large amount of
energy to enter the gaseous phase. This fact does not necessarily mean a higher
COP for water because; a higher pressure ratio is needed in the compressor. On
the other hand, higher energy gain during evaporation means in each cycle of
refrigerant water, a significant amount of energy can be removed from the
refrigeration compartment compared to the other refrigerants with the same mass
flow rate. Therefore, for a fixed refrigeration capacity, water works with a lower

mass flow rate compare to other common refrigerants.

 

 

 

390: T-s diagram . g .
340' 2 2 R718
290? _+ .
240 . [I [[[[[
H 190 [ ., R123.
2. 140» i WHY,“
" 90E fl ~
40) ._ “290 \. .\
-10:‘r/:,' - , E ~ .
:0: ER134a. 1 R22!

 

0 1 2 3 4 5 6 7 8 9 1 0
s [kl/k9°CI
Figure 8: Temperature-Entropy diagram for common refrigerants

In refrigeration cycles, Carnot standard cycle is considered an “ideal” cycle. If

a Carnot cycle is considered operating between evaporation and condensation

19

temperature, any deviation from ideality is considered irreversibility of the cycle.
With this assumption, the irreversibility of a simple refrigeration cycle comes from
two internal sources: desuperheating and throttling.

Figure 9 shows the irreversibility associated with desuperheating and
throttling for a typical refrigeration cycle. Based on this graph, the entropy

production during desuperheating process is

hb c
3 rod. = ‘(Sb‘s ) (2)
p 7bond. 0

 

, and the entropy production during throttling is

sprod. = Sr - 3d (3)

 

390

290

190

11°C]

90

 

 

 

 

-10 ----

 

 

 

 

 

-110

 

012 3 2r ‘5 ‘5 713010
SlkilkaK]
Figure 9: Entropy production during desuperheating and throttling for R718 and R134a

2.2.1 DESUPERHEATING AND THROTTLING IRREVERSIBILITY FACTORS
To compare irreversibility of different refrigerants, two factors are to be

introduced: desuperheating factor and throttling factor.

20

For desuperheating, two parameters contribute to irreversibility of a
refrigeration cycle. One is the inclination of constant pressure lines in a T-s
diagram and the other one is the entropy change during desuperheating. The
following is the derivation of constant pressure lines inclination in superheated
region for any refrigerant.

The differential form of the second law of thermodynamics is

dq
ds=—— 4
T ()

which indicates that the change in entropy is equal to the inverse of the

temperature integrated over the heat transferred.

Using the first law of thermodynamics de = dq -dw, the definition of work for a
gas dw = pdV, and the enthalpy definition h = e + pV, the differential change in

heat transferred can be written as
dg = dh - Vdp (5)
which shows an alternative way of presenting the first law of thermodynamics.

Equation of state for an ideal gas is
T = — (6)

For a constant pressure process based on equation 5, dq = dh. Also, the heat

transferred for a gas is defined as heat capacity of the gas times the change in

temperature.

dq = CdT (7)

21

Therefore,

dh = cpdr (8)

,where Cp is the heat capacity in constant pressure. Substituting 8 in 5, and its

result and 6 in 4, the following relation is obtained

dT dp
ds=C —-R— 9
p T p ()

Equation 9 can then be integration from state ”2" to state ”3" to give

I2 P2
As = Cp aveln——Rln— 10
( ) I3 P3 ( )

For an isobaric process, equation 10 simplifies to

T
As=<cp).v.In-T—§ (11)

Figure 10 shows a zoomed region around saturation vapor line for a refrigerant.

During an isobaric process from T2 to T3, the slope of constant pressure line is

tana = M (12)
As

Substituting equitation 11 to equitation 12, the following equitation is obtained

tana = IUTZ -InT3 (13)

(Cp )ave

 

22

 

/ const. Pressure Line

 

 

 

 

 

 

T1
vapor saturation line
m

Figure 10: Slope of a constant pressure line in Temperature-Entropy diagram

 

The numerator of equation (13 is the logarithmic average between temperatures

T2 to T3. Therefore,

L(Tz - T3)
p

tan(a) = (14)

Equation 14 shows the slope of constant pressure lines leaving the saturation
area to the superheated region for any refrigerant. Therefore, for a constant
temperature change from T2 to T3, the slope of constant pressure lines are
inversely related to Cp average between the two temperatures. Also, As for each
refrigerant determines the slope of saturated vapor line for a constant

temperature lift. For the purpose of comparison, a parameter Am has been

introduced as

AS

1 5
(Op )ave ( )

 

Arm =

A”), is a factor for desuperheating irreversibility comparison of different

refrigerants at a constant temperature lift. The higher the desuperheating

23

factor/Mm, the higher the desuperheating irreversibility associated with that

refrigerant is.

 

 

Ta
"9.
3 (g
3 as
.9 S
Fr? thronII'n "
é , Q = —g x 100 9/
30'? transfer %
S
{g-
Sthrothfing " Stransfer

 

 

 

Figure 11: Throttling irreversibility factor in a T-s diagram
During the throttling process, however, any change in entropy is identified as
irreversibility of the throttling process. As shown in Figure 11, the amount of
irreversibility has to be compared to the entropy transfer during the evaporation
process. For this reason, the throttling irreversibility factor Q is defined as

S .
Q=Mx1oo (16)

transfer

80, the higher the Q factor the higher the throttling irreversibility associated with

a specific refrigerant.

    
  

0.01
5.11 2.94 0.76 5.53 5.64 0.26

 

 

 

 

 

 

 

 

Table 2: Comparison of irreversibility and throttling factors for several
refrigerants for Tevap_.=5°C and Tcond..=35°C

24

Table 2 shows Am and Q factor for some selected refrigerants for a 30°C

temperature lift. As shown, water has the most desuperheating irreversibility
among the investigated refrigerants. On the other hand, R718 has the least
throttling irreversibility. This interesting fact shows that any improvement in the
compression process will significantly improve the efficiency of R718 cycles. A
simple theoretical calculation shows when R718 is used in multistage
compressors with intercooling, it achieves a significant COP increase.

Table 2 shows the desuperheating factor of 0 for R123. The reason is that R123
vapor saturation line has positive inclination. Therefore, when a refrigerant leaves
the evaporator and, in ideal case, expands isentropically, it still will be inside the
T-s dome. Therefore, there will be no desuperheating irreversibility for this

refrigerant.

2.3. COMPARISON IN A PRESSURE ENTHALPY DIAGRAM

One challenge when using water as a refrigerant is that for common
refrigeration conditions water has to work under vacuum. Figure 12 compares
water to the two most common refrigerants, R134a and R717 (ammonia), in a log
pressure- enthalpy diagram. In this figure, an operating condition of 5 °C
evaporation temperature and 30 °C temperature lift is considered. It is noticeable
that R718 cycle is operating in less than 1 atm pressure. Also, the enthalpy
change during evaporation for R718 is much more compared to the other two.

Although the specific enthalpy of R718 during evaporation is highest of all

among investigated refrigerants, its volume flow is still the highest. To compare

25

volume flow of investigated refrigerants, a non conventional T-Ah/v is shown in

 

 

 

 

 

Figure13.
100000
"[1111

10000 R134a ‘ _ .

1000 , I‘ y .
a Q 1atm
a 100 , ._
a / . ;

 

/[ 69 R718 cycle works at sub-ambient pressure I I
1 _ . ; , , '.L I , 3.;

4 .1

 

 

0.1

 

o 500 1000 1500 2000 2500 3000 3500
h [kl/k9]

Figure 12: Log pressure-Enthalpy diagram for R718, R717, and R134a

Based on Figure 13, R718 has the least cooling capacity per volume. In
another word, R718 has the highest specific volume flow, i.e. the volume of the
flow that must be compressed to obtain a unit cooling capacity.

Table 3 compares common refrigerants for a fixed typical evaporation and
condensation temperature. Among all, water requires the highest pressure ratio
for a fixed temperature lift, which results in a high temperature discharge for the
compressor. Therefore, in refrigeration cycles working with water as a refrigerant,

a multistage compressor is used with intercooling.

26

390
340
290
240
190
140
90
40
-10
-60
-1 10

T [°C]

 

0 0.02 0.04 0.06 0.08 0.1 0.12
Ahlv [kjlcm’]
Figure 13: Temperature-Enthalpy per specific volume diagram

 

 

 

Quantity R718 I R134a I R717 IR290] R22 1 R123
Evap. Temp [°C] 5
Cond. Temp [°C] 35

 

257 t — 1
Comp. Temp [°C] 178 0:: :2: 46 99 45 60 46

 

 

 

 

 

 

 

 

 

 

 

 

 

Specific Volume [m3/kg] 147 0.008 0.002 0.002 0.0008 0.0006
Mass Flow Rate
0.15 2.20 0.31 1.74 2.05 2.28
100 ton [kgs]
Compression Ratio“ 6.4 2.5 2.6 2.2 2.3 3.2
Pressure Lifi [Kpa] 4.75 537 835 667 545 897
COP Carnot 9.27
5.01 - ta
COP" one s g8 5.50 5.46 5.42 5.45 5.70
5.54 two-stage

 

 

 

 

 

‘Two-stage calculations are based on perfect intercooling and the same pressure ratio for each
stage.

"' COP calculations are for the same polytrcphic efficiency of 0.7.

Table 3: Comparison of different refrigerants in a typical evaporation and
condensation temperature

As shown in this table, when water is used with multistage compressors, its COP

is competitive to common refrigerants. Also, because compression part is the key

27

for improving refrigeration cycles working with water, any cycle modification
leading the use of innovative cycles or compressors will result in better

performance of water compare to other investigated refrigerants.

2.4. COEFFICIENT OF PERFORMANCE COMPARISON

2.4.1 COMPARING COP OF WATER (R718) AND R134a

R1343 is a very attractive refrigerant with zero ozone depletion and is the
refrigerant of choice in automotive air-conditioning. As shown in Figure 12, the
disadvantage of R1343 compare to other refrigerants is the low specific

refrigerating capacity.

 

-100 -60 -20 20 60 100 140 180 220 260 300
Te [°C]

Figure 14: COP isolines for R718 and R1343 refrigerants

Figure 14 shows the constant COP lines of water and R1343. In this graph,
the abscissa is the evaporation temperature and the ordinate is the temperature
lift. In this comparison, the same pclytropic efficiency is considered (assumed

0.7). The use of a pclytropic rather than isentropic compression efficiency

28

ensures that the overall performance of the compression process was not
influenced by the number of compression stages. Also, in these graphs the effect
of heat exchangers is reflected by moving the graphs of R1343 and R718
diagonally to the origin. The background comes from the benefit of using direct
heat exchangers in the case of R718. For water in most cases, an internally
direct heat exchangers with 1°C temperature difference plus external plate heat
exchangers with 1°C difference (equal to 2°C) are used. Whereas with R1343,
normally 5°C can be assumed for internal indirect heat exchangers. This gives
2°C or 5°C respectively for R718 and R1343 for the evaporator (T e) and again
2°C or 5°C respectively for R718 and R1343 for the condenser (T .m). In this graph
no subcooling or superheating is considered for R718. For R1343, 10°C

superheating and 5°C subcooling is considered (Table 4 and [27]).

 

Table 4: Marginal conditions for the performance comparison in Figure 14, taken
from Ref. [27]

In Figure 14 the maximum of each COP isoline is connected with a solid
yellow line. This line represents the best operating point for each refrigerant.
From this line it is evident that, in the case of water, the higher the evaporation

temperature, the higher COP for 3 fixed temperature lift will be. In this graph the

29

intersection of same isolines are connected with 3 dashed green line. This line
represents that for any evaporation temperature to the right of this line, the
choice of water as a refrigerant with a fixed temperature lift will result in 3 better
COP. The graph also shows that for relatively high temperature lift (greater than

60K) choice of water as a refrigerant represents better COPs.

2.4.2. COMPARING COP OF WATER (R718) AND AMMONIA (R717)
Ammonia is an excellent refrigerant with 3 high latent heat and excellent heat
transfer characteristics. However, because of its toxicity, it is often restricted to

applications in which the equipment is outdoors to allow natural dilution of any

leaks.

115
105
95
85
. 75
65
. 55

- 45
35
25
15

. . ‘ , 5
-100 -60 -20 20 60 100 140 180 220 260 300

Te [°C]
Figure 15: COP isolines for R718 and R717 refrigerants

T lift [°C]

 

Figure 15 shows the constant COP lines of water and ammonia. In this graph,
the abscissa is the evaporation temperature and the ordinate is the temperature

lift. In this comparison, like the pervious one, the same pclytropic efficiency is

30

considered (assumed 0.7). In these graphs the effect of heat exchangers are
reflected as described in the pervious comparison. In this graph, the solid line
connecting the maximum of COP isolines of ammonia covers a wider
evaporation temperature range than that of R1343. However, in this case for low
evaporation temperatures and low temperature lifts water performs better in

terms of COP.

31

CHAPTER 3: IMPLEMENTATION OF A CONDENSING WAVE ROTOR

3.1. WAVE DEVICES

A gas in 3 supersonic flow can be compresses in three methods: Isentropic,
Prandtl-Meyer, and shockwave. The method of compression of 3 gas results in
different temperatures and densities for 3 given pressure ratio, which can be
analytically calculated for 3 non-reacting gas. A shock wave compression results
in a loss of total enthalpy, meaning that it is a less efficient method of
compressing gases for some purposes, for instance in the intake of 3 scramjet.
However, by generating shock waves in appropriate geometries, unsteady wave
machines can transfer the energy of a high-pressure fluid directly to another low-
pressure fluid without using mechanical components, such as pistons or vaned
impellers. This is one of the reasons that the potential for using wave devices in
thermodynamic cycles for power generation, propulsion, and refrigeration has
attracted the attention of researchers since the early twentieth century. Shock
tubes, shock tunnels, pressure exchangers, pulse combustors, pulse detonation
engines, and wave rotors are among the best-known wave devices developed
thus far. These devices represent applications of classical nonsteady, one-
dimensional compressible flow theory. It is well known, but not yet widely
employed, that time-dependent flow devices can generate much greater pressure
rises than those obtained in steady-state flow devices [40-42].

Within the family of wave devices, wave rotors have demonstrated a
favorable potential for reaching the ultrahigh performance targets of power

systems and for lowering their cost. A wave rotor is a dynamic pressure

32

exchanger that exchanges energy from a high pressure fluid to a low pressure
fluid via the means of moving shock waves that are generated in an appropriate
geometry by exposing the ends of flow channels alternatively to high pressure

and low pressure ports (Figure 16).

 

Figure 16: Schematic configuration of a typical wave machine [43]

Much research has been devoted to the progress and development of wave
rotor technology, especially in the past few decades. Most of the work in the US.
expands from the early efforts of the General Electric Company (GE), General
Power Corporation (GPC), and Rolls Royce [44,45] to the recent activities of the
wave rotor team at NASA Glenn Research Center (GRC) collaborated by the
US. Army Research Laboratory (ARL) and Rolls-Royce Allison [46-55].

In recent years, Turbomachinery Laboratory of Michigan State University has
been actively pursuing research in different propulsion and power generation
purposes. Research teams at Warsaw University of Technology in Poland,
Purdue School of Engineering and Technology in Indianapolis (lUPUl), and
University of Tokyo are other examples of active research in the area of unsteady

wave devices, specifically the wave rotors [56].

33

3.2. WAVE ROTOR

Utilization of unsteady flow devices to enhance the performance of
thermodynamic cycles has been a major branch of study since the early
twentieth century. The basic concept underlying these devices is the transfer of
energy with shock waves. By generating compression waves in appropriate
geometries, unsteady wave machines can transfer the energy of a high-pressure
fluid directly to another low-pressure fluid. Thus, a pressure rise can be obtained

without using mechanical components such as pistons or vaned impellers. [57].
3

 

 

/ 77 _] 00%
2

Difl‘

M MSM M M
1 2 1
——> —-> —> —>

(I) (2) 1) (2)

 

Moving shock wave Diffuser

ID2I I"1

1.5 '

 

 

 

 

Figure 17: Comparison of pressure gain (static pressure ratio) of moving shock
and steady flow isentropic diffuser for y=1.4 [42]

For instance as shown in Figure 17, for the same change from a given inlet
Mach number M1 to a given outlet Mach number M2, the static pressure ratio
p2/p1 across a single shock wave moving in a frictionless channel (solid line) is

always much greater than that obtained by an isentropic deceleration in a 100%

34

efficient diffuser (doted line). However, it must be noted that friction effects
always exist, and hence pressure gains are actually less than those predicted by
Figure 17. It is also useful to compare the shock wave isentropic efficiency
(773m).) with the diffuser efficiency (770,”). Figure 18 shows variations of ”Shock
(solid line) and 770m (dashed line) as functions of the pressure gain p2/p1 obtained
by a moving shock wave in 3 frictionless channel and by 3 diffuser, respectively.
770,-” is calculated for different values of total pressure drop across the diffuser
expressed by p22/pu. Similar flow friction-effects would lower the efficiency of
wave devices and reduces the efficiency advantage of such devices, which is not

shown in Figure 18 [42].

 

 

 

 

 

 

 

1
>. 0.8
5
E I /771)ijf (”2) V -1 6
m j I/ P1 5
g 0.6 h, I, ”Diff: Y—1 II 0.6 g
'3 ,1, [P11 P2] v _1 E
9 I, Pr2 P1 3
E 0.4 II _1 . 0.4 g
0 I Y— a
g I (i—Z) V 1
1
': rIshock =
“’ Y+1 P2
0.2 —+— ' 0.2
P2 Y-1 P1 1
P1 1+Y+1p2
Y-1P1
0 l l l I o
1 1.5 2 2.5 3 3.5 4
Wm

Figure 18: Shock wave and diffuser isentropic efficiencies as a function of the
static pressure ratio [42]

35

The comparison in Figure 18 reveals that for the same pressure gain p2/p1,
shock compression efficiency may far exceed the efficiency obtained by a
diffuser. Thus, it is seen that for a pressure recovery of up to p2/p1=2.2, 773nm is
greater about 93% and hence it is always greater than that of a usual diffuser.
Therefore, by substituting a diffuser with an unsteady flow device utilizing shock
waves, a gain in cycle performance can be expected.

There are at least two classes of wave machines: rotating wave machines
and non rotating wave machines. Dynamic pressure exchangers and wave rotors
are two types of rotating wave devices developed to reach the high performance
targets of thermodynamic cycles. In Europe, however, both terms were often
used interchangeably [58]. The first available literature on dynamic pressure
exchangers utilizing pressure waves is a patent in 1929 filed by Burghard [59].
This patent has introduced the concept of the dynamic pressure exchanger. The
essential feature of these devices is an array of several channels arranged
around the axis of a cylindrical rotor. The assembly rotates between two end
plates, each of which has a few ports or manifolds that control the fluid flow
through the channels. Through rotation, the channel ends are periodically
exposed to the steady state ports located on the end plates causing the initiation
of compression and expansion waves within the passages. Therefore, unlike a
steady-flow turbomachine component, which either compresses or expands the
gas, both compression and expansion are accomplished within a single
component. A reversed design with stationary rotor and rotating ports is also

possible [60]. Such a configuration may be preferred for laboratory investigations

36

because it easily enables flow measurement in the channels where the important
dynamic interactions take place. However, this arrangement rarely seems to be

convenient for commercial purposes [61].

   
 
 

Flow
Passages

  

From

2
End Plate To Combuster

Figure 19: Schematic configuration of a typical wave machine

Figure 19 schematically represents most of the rotating wave devices. Each
end plate contains a few ports. To minimize leakage, in practice, there is only a
very small gap between the end plates and the rotor and the end plates with
sealing material may contact the rotor. The rotor may be gear or belt driven [61]
or preferably direct driven by an electrical motor (not shown in the picture). The
power required to keep the rotor at a correctly designed speed is very small, it is
almost zero [62]. It is just an amount necessary to overcome rotor windage and
the friction in the bearing and contact sealing if used. Alternatively, rotors can be
made self-driving. This configuration called the “free running rotor" can drive itself
by angling the blades against the direction of rotation [63]. In this case, the
momentum of the inflow or outflow rotates the rotor.

In dynamic pressure exchanger and wave rotors two basic fluid-exchange

processes usually happen at least once per revolution of the rotor: the high

37

pressure process (charging process) and the low pressure process (scavenge
process). In the high pressure process, compression waves transfer the energy
directly from a fluid at a higher pressure to another fluid at a lower pressure. The
low pressure process employs expansion waves to scavenge the flow from the
rotor channels. Ingestion of the fresh cold fluid into the rotor channels is also

performed in this process [42].

Combustion Chamber

 
   

0

4

Figure 20: A schematic of a gas turbine topped by a 4-port wave rotor [42]

 

 

Advantageous of using wave rotor can be summarized as:

o Potentially have higher efficiency than turbomachines

Higher overall engine pressure ratio
. More compact systems as they combine compressor and turbine stage in one

device

Lower rotational speed compared to turbomachines, which results in low

material stresses

Simpler design and geometry compare to turbomachines. Therefore, lower

manufacturing costs.

38

. Channels are less prone to erosion damage than the blades of turbomachines.
This is mainly because of the lower velocity of the working fluid in the
channels, which is about one-third of values within turbomachines [61].

. Self-cooling capabilities. Wave rotors are naturally cooled by the fresh cold
fluid ingested by the rotor. Therefore, applied to a heat engine, the rotor
channels pass through cool air and high-pressure and high-temperature gas
flow in each rotor revolution. Therefore, the rotor material temperature is
always maintained between the temperature of the cool air which is being

compressed and the hot gas which is being expanded.
3.3. APPLICATION OF WAVE DEVICES IN REFRIGERATION CYCLES

As discussed above, refrigeration cycles working with water as a refrigerant
require bulky multistage compressors; therefore, making them not suitable for
small scale applications such as home size refrigeration. On the other hand,
application of wave rotors in the gas turbine cycles and their significant energy

savings introduce a potential application in cycles involving phase change.

Increasing the pressure difference in liquid state requires a simple pump;
however, when the substance is in vapor phase, a high technology compressor is
required to provide a high pressure ratio. Therefore, for instance, in the R718
cycle, if we could provide a portion of pressure ratio required in vapor phase in

the liquid state, the size and the cost should decrease.

39

Base on the above idea and ongoing research on wave devices and
refrigeration cycles at Michigan State University, a novel idea of Condensing
Wave Rotor sparked which later, lead to a US patent application by Norbert
Mueller.

Adding a wave rotor to 3 R718 cycle enables greater temperature lift or
reduces the compressor pressure ratio, which is crucial for the R718 chiller
technology, where the stage pressure ratio is very much limited by the
thermodynamic properties of water vapor. Although invention of Condensing
Wave Rotor was started in R718 cycles, its application can be investigated for
other refrigeration cycles or, even beyond that, any cycle involving phase change
and compression of a gas.

The present study demonstrates the enhancement of a turbocompression
refrigeration cycle that uses water as refrigerant (R718) by utilizing a novel 3-p0rt
condensing wave rotor. This device is meant to be designed such that its
existence in the cycle compensates for one stage of compressor. At the same
time, the condensation of the refrigerant occurs inside its channels. Therefore,
overall size of the refrigeration equipment is expected to be reduced significantly.

In the following chapters, beside investigation of R718 cycles and wave
rotors, 3 detailed study of condensing wave rotors, and its design, and analytical

study is presented.

3.4. CONDENSING WAVE ROTOR
The novel idea of condensing wave rotor springs from extensive research for

implementation of conventional wave rotor in gas turbines. Wave rotors appear in

40

different configurations and different number of inlet and outlet ports. As
described in previous chapters, in R718 refrigeration cycles, the key component
is the compressor. To achieve better COPs and smaller R718 units, one solution
could be implementation of a wave rotor in R718 units. For this purpose, a 3-port
wave rotor is considered. Concerning the pressure levels, 3-port wave rotors’

function is similar to pressure-equalizers’ [61] like the one depicted in Figure 21.

6(hifll'Pmsum)—' \

 

”M.""

‘ “5”” —b 3 (medium-pressure)
“‘o‘

‘u

  
  

  

  

"N...

v-
N...

  
  

2 (low-pressure) —>

 

Figure 21: Dynamic-pressure equalizer

The phase change of the fluid inside the wave rotor in 3 R718 refrigeration
application is a major difference to the operation of a wave rotor in a gas turbine
cycle. Additionally, here the low pressure fluid is at higher temperature than the
high pressure fluid. In this innovative design, condensation of vapor occurs inside
the wave rotor; therefore, it is called condensing wave rotor. A schematic of a 3-
port condensing wave rotor is depicted in Figure 22. The condensing wave rotor
provides energy exchange in liquid state which is much more economical and
easier than vapor phase. The liquid leaving the device is pumped to a higher
pressure level, and then, is introduced suddenly to a low pressure vapor phase in

the channels. The details of process are discussed in future chapters.

41

R718 cycle is the first application researched for condensing wave rotor. This
novel idea has a potential to be implemented in any application with phase
change, where providing the required pressure ratio is difficult. Also, water is the
first refrigerant researched for condensing wave rotor fluid. Condensing wave
rotor has the potential of being used on other refrigeration cycles such as CO2
refrigeration cycles or even R1343.

Simply, because of the novelty of the condensing wave rotor, the author only
focuses on application of condensing wave rotors in the refrigeration cycles, and

only for refrigeration cycles using water as a refrigerant.

42

a

E can;
2:»de £235.

033 Em

C 552,
2:82.. 5....

      

8V cons)
05323 33

582.3 coqm>

Figure 22: Schematic of a 3-port condensing wave rotor

43

CHAPTER 4: ANALYSIS OF CONDENSING WAVE ROTORS

4.1. INTERNAL: TRANSPORT AND PHASE CHANGE PHENOMENA
A schematic thermodynamic model of a R718 cycle using a 3-port

condensing wave rotor is depicted in Figure 23.

 

 

 
   

 

 

    

 

ma mm poumo

. .4' . 3 .. ..‘
7 ' ' . '«7' \.
_ 3:}; r.
'> 31 r‘ .
. “ . ‘i
r_ '1 ‘ -
L I I o 1 .
..f

Cooling water cycle

  

  

-. .. -' at?

'- . ‘ t‘ .‘.'

.. ‘ - .

'5...‘ ‘ - '- s:

. is r'. . .‘ ‘.-'

£31..” 3.3,?

; r~., .‘ 4:

. - v'

' .ff :1,‘. .‘. .‘o‘f

'o'a‘x- q .‘.;I"'..

¥
7

w” +Wm throttling valve

@ cooling bower

  

it

Figure 23: Schematic of the thermodynamic model of a R718 chiller unit
enhanced by a 3-port condensing wave rotor (CWR) substituting for the
condenser and for one stage of compressor

While Figure 22 shows a schematic of a 3-port condensing wave rotor, Figure
24 schematically shows the regions modeled for a channel during compression
and condensation. The following explanation follows the points (states)
introduced in Figure 23. Coming from the turbocompressor (2), the superheated
vapor flows continuously through a vapor collector (shown in Figure 22) to the
inlet port of the wave rotor located at one of the two stationary end plates. By
rotating the wave rotor between the two end plates, the wave rotor channels are

opened to the port and filled with the incoming superheated vapor.

 
    
  

    

axial length

mm; '

:9 fiofiéfifio

2 -.'<~’_1Lv <

- Elm: mwmfivnfiw ' ' ., .

HIIXW‘KSOVCWA‘WV Am’gfifi' - - Lre- -<
4 i "

 

 

‘. ‘4 ‘-‘<~.<:¢:~s.’~=:&?1'3'.n_: :cn'g -

“dkfll’léw

 

m
v

Figure 24: a) Schematic wave and phase-change diagram for the 3-port
condensing wave rotor (high-pressure part) b) A magnified channel showing the
regions modeled during compression and condensation.

Region (a) in Figure 24 is the state after the filling process is completed. After
further rotation. the channels meet the second-inlet port (6) through which the
high-pressure low-temperature water (9) comes in and is exposed to the low-
pressure high-temperature superheated vapor in region (a). Because of the
sudden pressure drop (from p6 to p2), all the heat cannot be contained in the
incoming water as sensible heat and the heat surplus is transformed into latent
heat of vaporization. This is the so-called flash evaporation or flashing
phenomenon [57, 65]. Therefore, one portion of the incoming water suddenly
vaporizes (c) and the remaining part cools down (d). The frontal area of the
saturated vapor (c) generated by the flash evaporation is called the contact

interface and acts like a fast-moving piston. It causes a shock wave triggered

45

from the leading edge of the inlet port traveling through the superheated low-
pressure vapor, which exists inside the channel (a). The shock wave travels with
supersonic speed (VM) faster than the contact interface Mum). Therefore,
the trajectory of the shock wave (solid line in Figure 24) has a smaller slope than
the incoming water and the contact interface of the generated vapor (dashed
line). Behind the moving shock wave (b) the temperature is increased from T2 to
T2» and the pressure is increased from p; to p2: = p; because of the shock
compression. The latter is a design decision similar to a tuning condition. With it,
the pressure at the inlet port p6 is set to an appropriate value that generates the
pressure ratio pa/pz required to trigger the desired shock wave. The superheated
vapor will be condensed at pressure p3 . This shows that the fluid in its liquid
state serves as a “work capacitor” storing pump work to release it during its
expansion in the wave rotor channels for the simultaneous vapor compression.
Therefore, in the enhanced system the pump in the cooling water cycle not only
has to provide the work necessary to overcome the pressure loss in the heat
rejecter cycle pr, but also the work necessary for the shock wave compression
in the wave rotor channels Wpc. The pressure behind the shock wave (b) is
imposed on the vapor generated by the flash evaporation (c). It is the pressure at
the water surface and the equilibrium pressure at which the evaporation decays
p(c) = p(b) = p3 . Hence, both generated vapor and the cooled water obtain the
saturation temperature T3 = Tsat (p3).

Because of the direct contact of the superheated compressed vapor (b) with

the cold incoming water (a), the superheated vapor is desuperheated and its heat

46

is transferred (f) to the incoming water. This continues until the equilibrium
temperature T3 is achieved in region (b) and the superheated vapor is changed
to saturated vapor. Subsequently, the incoming water compresses the saturated
vapor further and condenses it, while the latent heat is transferred to the
incoming water (9). The water, which is nearly a fully condensed two-phase
vapor with a typical quality of 0.005, is scavenged through the only outlet of the
wave rotor (3). The scavenging process may be supported by gravity and pump

power.

axial length

 

 

 

 

 

 

 

 

Figure 25: Schematic wave and phase-change diagram for the 3-port condensing
wave rotor including the discharge process

Figure 25 shows the discharge process for a design of condensing wave
rotor. In this design, the only outlet port in opened exactly at the same time as
the low pressure port is opened. Therefore, column of water will flow downward
because of gravitational force to the water mass inside the channel and low

pressure water vapor coming from compressor takes the emptied space. The

47

outlet port is kept opened until the liquid water is fully scavenged. Then, the cycle

repeats.
4.1.1 STUDY OF FLOW INSIDE A CHANNEL OF A CONDENSING WAVE ROTOR

Contact surface (interface between the high
pressure and bw pressure gas) acting as a
piston

(a) @g;  -@-2~ 0

‘L7 H Vinrerface" ————> VSI'OC"

 

,m, T,” 13"" , _
I}, Pry/’3 9}“ p3 pf'g'ps 'P2 12' '1 1P2 4P2 72

\

Figure 26: Regions modeled during compression and condensation as described
in section 4.1

 

 

 

 

Normal shock wave

For simplicity, only a single channel of condensing wave rotor is considered in
this section. Based on the flow regimes described in section 4.1, the shock wave
calculations are presented, leading to the “characteristic equation”, which
determines the inlet pressure required for incoming high pressure fluid so that the

outlet fluid is saturated liquid in a prescribed outlet pressure.

4.1.1.1. TEMPERATURE INCREASE ACROSS THE NORMAL SHOCK WAVE

Consider p3. T2, and p2 are known variables, and Tzvis a known variable. p3 is
the outlet pressure of the condensing wave rotor. T2 and p2 are the temperature
and the pressure of the low pressure port; on the word, the outlet state of the

compressor.

48

Defining the Mach number as

M = Vshock
32

(17)

where a2 is the speed of sound in region a, the Mach number is related to the

pressure ratio across the wave [67]

31:—21—(M2—1)+1
P2 7+1

From equation 18

 

M =\/lf_1(fli_1)+1
2? P2

The relationship between the temperatures across a normal shock is [67]

7-1 2 27 2
1 -———M -——-———M -—1
T2. -( + 2 Xy-1 )

T2 (7+1)2 M2
2(7—1)

where T2» is the temperature in region b and can be determined as

7-1 2 27 2_
(1+—2-M X73,“ 1)
(1+1)2M2
2(7-1)

 

 

Using equation 19 and 21, the following equation is derived

49

(18)

(19)

(20)

(21)

(1+7'1[-7——2‘;1(p——3 -1)+1])(2——1_[“—1(p—3 1)+1]— 1)

 

 

T2-=T2 2 ”2 2 2 ”2 (22)
(7+1) [72+1(P3_1)+1]
2(Y 1) 27 P2

Equation 22 shows that by knowing the pressure of outlet p3. temperature T2,

and pressure p2 of low pressure inlet, the temperature in region (b) T2. can be
obtained.

4.1.1.2. INTERFACE VELOCITY ACROSS THE NORMAL SHOCK WAVE
Again, p3, T2, and p2 are known variables. The following derivation is to
calculate the the interface velocity Vimmoe across the normal shock wave.

The relationship for the density across a normal shock wave is [67]

3.2:?” M2 (23)
92 2 ”Lg—1M2

Substituting equation 19 in the above equation gives

 

 

 

 

\
. i [m(p—3—1)+1]
2 -
p2 1+1_1[lfl(2§__1)+1]
( 2 2V p2 )
or
. 1+(y—+1)B§
2;: Y‘1 DZ (25)
92 Lfl+P_3
Y-1 P2

50

Assuming V2 is zero (low pressure vapor is stagnant when reaching the high

pressure port), the continuity equation between regions a and bis

P2Vshock = PZ (Vshock " Vinterface)

Therefore,

 

£2 = Vshock = M
p 2 Vshock "' Vinterface M _ vinterface
32

from which VmMm/a2 is determined as

Vinterface = M(1—£2-)
32 92

Now, applying equation 19 and equation 24 to the above equation gives

 

 

 

 

 

{ 1:1,)”;
v.
mterface=\[y_+1(_P_3__1)+11_ y—1 p2
a2 27 P2 1+(Y_+1)B§
7-1 P2
or
P3
2—-1
vinterface =Jfl£§+y—1 (p2 )
a? 27 p2 2* (1+1)"—3+(y-1)
P2
or

Vinterface =iifl-1X 2 )0.5
82 )5 ”2 (1+1)-g1+<y-1)
2

 

 

51

(26)

(27)

(28)

(29)

(30)

(31)

Speed of sound in region a is az = ,ly.R.T2 ; therefore,

ZRT
Vinterface = (2—3 ‘1)(——2—)OI5 (32)

2 n+n3§+c-n
P2

Therefore, knowing the pressure of outlet p3, temperature T2 and pressure p2 of

low-pressure inlet, the interface velocity wnterface can be calculated.

4.1.1.3. CHARACTERISTIC EQUATION
The speed of interface equals the speed of incoming water V,_ plus the speed
of evaporation V9,,

_ -2:
Interface "' vL +Vev — 96A + pgat—QA (33)

 

where p3“ is the density of region 0 and A is the cross sectional area of the

channel. Saury [57] has introduced the flash evaporation rate coefficient J as

mev

J=— 34
Nm—m) .()

Penetration depth

 

7

Time
Figure 27: Flash evaporation photography by Miyatake [65]

52

Miyatake [65] has noted that J is not time dependent and has determined the

value of J for several conditions. For other conditions J must be interpolated,

extrapolated or determined through experiment. Thus, knowing J, may can be

 

 

 

 

 

 

 

 

 

 

    

 

 

 

 

obtained by
mev = JA(P6 _ P2) (35)
2RT m 9‘3"“
P6 =P2 +(fi—1)(—'p2_ 0'5 -——2 'LJ—
”2 <v+n—§+w-n ”5
P2
VL m6 , 1’63; _ = A./(p6 — p2)
L 961‘! g" I - 2 9V pgm‘gA
” i '. .'+::1:$ 1 '9'! E'r‘ ’
91 , ' ”5: 33%:
\ 3 \
' + 1
j‘Vs:aZ X2— (P_3__+1) 1 l \ i
v p2 ‘v _ _ 2er Q5,
therface— (_P3 1)( ———‘—) .‘

; therefore,

 

”2 (7+n9§+w—n
P2

Figure 28: Condensing wave rotor wave diagram

fl... iAl_P§_£2_) = (£1 -1)( _ 2RTZ )05

(36)
96A 93"” ”2 (v+1)P-‘1+(v-1)
P2

or solving for p6

53

sat-g

2RT I176 p
p6=p2+(”—2—1)( p 2 )°-5— A 2,
”2 (v+1);2+<v-1) “6
_ 2 _

 

 

(37)

 

 

The above equation is the characteristic equation of a condensing wave rotor.
In the above equation knowing the exit temp T3“! , the density p?“9 is known.

Therefore, knowing the conditions of the gas inside the channel and the exit
conditions, the high pressure port mass flow rate file may be a design factor for

determination of its pressure p6.

4.2. EXTERNAL: THERMODYNAMIC CYCLE

The low pressure inlet and the outlet pressure for the condensing wave rotor
was considered as known variables for the calculation of characteristic equation.
The thermodynamic states of the low pressure inlet port and the outlet port can
be calculated using the external thermodynamic analysis of the cycle.

The schematic pressure-enthalpy (p-h) and temperature-entropy (T -s)
diagrams of both the baseline and the wave-rotor-enhanced cycle are depicted
in Figure 29 and Figure 30 respectively. Both cycles start at the outlet of the
evaporator (state 1), where the vapor is saturated. State 2b represents the
compressor outlet of the baseline cycle, whereas state 2 is the compressor outlet
of the wave-rotor-enhanced cycle, which allows using a compressor with a lower

pressure ratio.

- - - - - - l-2b-3b-Sb-tSb-7b-3b-4b baseline cycle
-—-—- l-2-2’-3 -5-6-3-4 wave rotor enhanced cycle

 

Pressure

 

 

 

 

Enthalpy

Figure 29: Schematic p-h diagram of a R718 baseline cycle and enhanced cycle
with a 3-port condensing wave rotor.

State 2' is an intermediate state inside the wave rotor channels that
corresponds to the flow properties in region (b) right after the shock wave. The
slope between states 2 and 2' is greater than that between states 1 and 2.,
because the shock compression typically occurs with a higher efficiency [4042].
Still inside the wave rotor channel, the superheated vapor is desuperheated to
the equilibrium temperature T3 (2'—+3). State 3 is actually much closer to the

liquid region than shown in Figure 29 and Figure 30 because the mass flow rate

of the cooling water cycle (5..) is much greater than that of the core cycle (5.2).
Knowing this, it becomes clear why the distances between states 3, 5, and 6 are
magnified in this schematic diagram. The expansion process (6-43) releases the
energy consumed by the compression process of the vapor (2—+2') all within the

wave rotor channels.

55

- - - - - - l'zb'3b‘5b'6b’7b'3b‘4b baseline cycle
l-2-2’-3-5-6-3-4 wave rotor enhanced cycle

 

2’

Temperature ,’
I

 

 

 

 

 

 

 

 

Entropy

Figure 30: Schematic T-s diagram of 3 R718 baseline cycle and enhanced cycle
with a 3-port condensing wave rotor

Coming from the only outlet port of the wave rotor (state 3), the flow diverges.
The small fraction used as refrigerant is directed to the expansion valve and is
expanded in a constant enthalpy process (3—>4), while largest part of the flow out
of the wave rotor is pumped (3—+5), providing the energy for the vapor
compression in the wave rotor Wpc and compensating for the pressure loss in the
heat rejecter and associated piping pr. In this case, an inducer pump may be
used to avoid cavitations problems. Then, the fluid goes into the heat rejecter
(cooling-tower or similar) where it cools off (5—+6). In the enhanced cycle, the
inlet port of condensing wave rotor can be viewed as throttling; therefore, a
separate expansion valve is not shown. In difference to the insenthalpic
expansion (throttling), in the condensing wave rotor most of the pressure is
recovered immediately for compression of the refrigerant gas in the channels of a

condensing wave rotor. The expansion in the condensing wave rotor can be

56

viewed as a turbine expansion in which the work necessary for refrigerant

compression is extracted.

 

 

 

 

Figure 31: Novel compact R718 water chiller with integration of a condensing
wave rotor

Figure 31 pictures one possible integration of a condensing wave rotor in a
R718 cycle. This innovative design unifies a significant part of the compression
with the desuperheating and condensation of the refrigerant vapor in a compact
dynamic unit. Such innovative designs of R718 chillers with integration of
condensing wave rotors will result in manufacturable, easily scalable (3-300 kW),

and high-efficient refrigeration cycles.

57

4.3. SIZE COMPARISON

Using a 3-port condensing wave rotor in a water refrigeration cycle can
improve the coefficient of performance of R718 units [64] while reducing their
size and cost. lts successful implementation may replace three subsystems: the
intercooler, one compressor stage, and the condenser. With conservative
measures as shown in Figure 32, this may reduce the overall size of the R718
unit to nearly 50%, since the volume of these three subsystems reduces down to

about one-tenth of the current size [27].

/

 

= 8.5m

 

 

Condenser Evaporator 32m

.rossudtuog)
Io 3381s Pumas
.roloooaarnl
.rosseJdurog
10 0381s mu

 

 

 

 

 

 

 

 

 

 

 

Condenser

.rosseadtuog
1° 33318 plums
g \\\\\4
II
II
N
3

\

 

 

 

 

 

=1,“ —-l

 

 

=4.7m

 

 

Figure 32: Size reduction by combining intercooler, second stage compression,
and condensation into a condensing wave rotor (CWR)

It is noted that the wave rotor length is a design parameter that adjusts with
rotational speed and the speed of sound within the fluid. The wave rotor diameter

is governed by the volume flow rate of the precompressed vapor out of first

58

compressor stage into the condensing wave rotor and the number of wave cycles

per revolution.

59

5. PERFORMANCE EVALUATION OF CONDENSING WAVE ROTOR
INTEGRATED IN A R718 CYCLE

A computer code based on the thermodynamic model described above is
generated for performance evaluation of R718 refrigeration cycles enhanced with
3-port condensing wave rotors. The evaporator temperature T1 and heat rejecter
temperature T3 are commonly fixed by the application. The objective is to get the
highest increase in coefficient of performance (COPgaln) compared to the

baseline cycle. Independent design parameters are the mass flow ratio (K =

")6 / n12 ), which relates the mass flow of the cooling cycle to the mass flow of the

core cycle, and the pressure ratio of the wave rotor (PRw = p3 lp2).
Additional assumptions considered in the thermodynamic model are as follows:
a For comparison of baseline and enhanced cycles, the evaporator and

condenser inlet temperatures are considered the same (T1 = T13 and T3 =

T33).

Temperature difference across the heat rejecter is kept constant (T6 - T5 = 3

K).

Pressure drop in heat exchanger, evaporator, and pipes is neglected.

The condenser and evaporator outlet states are fully saturated.

The same pclytropic compressor efficiency is used for baseline and enhanced
cycles. Its value of 0.72 is obtained by assuming an isentropic efficiency of
0.7 for a compressor with a pressure ratio of 2.

The superheated vapor is considered as an ideal gas (v = 1.33).

60

- One-dimensional gas-dynamic shock wave equations are used to calculate the
flow properties across the moving normal shock wave. Reflected shock
waves are not considered.

0 The hydraulic efficiency of the pump is 0.9.

0 Liquid water is considered as incompressible.

5.1. THE BASELINE PERFORMANCE CALCULATION

In the ideal vapor-compression refrigeration cycle shown in Figure 4,
refrigerant from the evaporator flows into the compressor as a saturated vapor,
and then it discharges into the condenser as a superheated vapor. The saturated
liquid refrigerant at the condenser outlet returns to the evaporator through the

expansion valve and then cycle repeats.

 

 

 

    

.";-Tf’fff“_fflfibquataforthe _. aséllnegxgle :.
Saturation temperature of the T
evaporator 9
Saturation temperature of the T
condenser c

 

Table 5: Input data for the baseline cycle analysis.

As shown in Table 5, saturation temperatures at the evaporator and the
condenser are the input data for this analysis. To obtain the COP of the cycle,
the thermodynamic states at each location are determined sequentially as

follows:

Compressor Inlet (State 1).

P1 = Psat(Te) (38)
h1 = hsat (To)

61

Condenser Outlet (State 33).

Tab = To
par) = Psat (Tc) (39)
h3b = hsat (Tc)

Compressor Outlet (State 23). The cycle overall pressure ratio is calculated by

P1

and the enthalpy change across the compressor, assuming an average specific

heat, is obtained by

11c

Ah = [i]cpr1[(nb)ll‘1)’7] (41 )

where compressor isentropic efficiency ’10 is calculated by assuming a pclytropic
efficiency of 0.7. Therefore, thermodynamic properties of the compressor outlet

are
h2b = h1 + Ah
P2b = P3b (42)
T2b = T(P2b+h2b)
Expansion Valve Outlet (state 43).

h4b = h3b

P4b = P1 (43)

62

T4b = T(P4b1h4b)
By definition COP is the ratio between the processed heat at the evaporator
(qL = h1 - h4b) to the work consumed by the compressor (wc = h2b — h1b)

We

5.2. THE WAVE ROTOR ENHANCED CYCLE PERFORMANCE CALCULATION

As shown in Figure 23 in the enhanced cycle, the superheated vapor leaving
the compressor discharges into the wave rotor. The pressure ratio of the
compressor is less than that of the baseline engine. After the compression of the
superheated vapor in the wave rotor, one portion of the almost saturated water at
the wave rotor exit (3) goes to the heat exchanger, while the other portion returns

to the evaporator through the expansion valve.

 

 

 

. “Imutdatarorenhancedcycle " 7" ‘7
Saturation temperature of the T
evaporator 9
Saturation temperature of the T
condenser 0
Temperature drop in the condenser ATC
Mass flow ratio between chilled water k
cycle and the cooling water cycle
Pressure ratio of the baseline cycle 113
Pressure ratio of the wave rotor PRW
Hydraulic efficiency of the pump 77p

 

 

 

Table 6: Input data for the enhanced cycle analysis

63

The input data for analysis of the enhanced cycle are given in Table 6. To
obtain the COP of the enhanced cycle, the thermodynamic states at each
location can be obtained sequentially as follows:

Compressor Inlet (State 1). The compressor inlet condition at state 1 is the
same as the baseline cycle.

Wave Rotor Outlet (State 3). The wave rotor outlet flow is in the saturation

region, very close to the saturated liquid line. Therefore,

(45)
p3 = Psat (Tc)

Considering a control volume around the wave rotor, the conservation of mass

law gives

m2+ m5 = [173 (46)

and conservation of energy implies
h2m2+ ham6 = h3 a). (47)

Using the definition of mass flow ratio (K = ms/mz ), Eqs. (46) and (47) can be

combined as
h -(-1—](h +h K) (43)
3 1 K 2 6

The enthalpy at state 6 will be calculated later.

Evaporator Inlet (State 4).

P4 =P1

“4 = ’73 (49)

T4 = T(P4+h4)
and the quality of liquid is calculated by

X4 = h4 ‘hsat (Te) (50)
hsat (Te ) ‘hsat (Tc )

Compressor Outlet (State 2). The compressor exit pressure is calculated by

IIb
=_ 51

and Equation (41) is used to calculate the enthalpy at the compressor exit,

substituting the new value of compressor exit pressure. Therefore,

h2 = ’71 + Ah
(52)
T2 = T(P2.h2)

Shock Wave Compression (State 2'). As described above, because of the

sudden pressure drop from p6 to p2, flash evaporation generates a shock wave

triggered from the leading edge of the inlet port traveling through the

superheated low-pressure vapor, which exists inside the channel. Therefore, the

temperature is increased from T2 to Tzc and the pressure is increased from p2 to
p2- =p3. Using moving normal shock relations, temperature increase is

calculated by [66]

65

'&,v_+1
T.=T221L2_7_‘_1_

 

 

(53)
r-1P2_
and
P2' = P3
(54)
(12' = h(T2+ .va )

Pump Outlet (State 5). As stated above, the pressure (p5) provided by the

pump in the cooling water cycle is used to generate the pressure ratio

p6 / p2 required to trigger the desired shock wave. Therefore,

Ap=<h2- —h2)"75 (55)

where liquid water is considered incompressible; therefore. p5 = 03 = p(T3,p3).
Thus,

P5 = P3 + AP (56)
and the enthalpy increase by the pump is
Ah = i (57)
P571!)
where 11,, is hydraulic efficiency of the pump. Therefore,
h5 = ’73 + Ah
(58)

66

Ts = T(P5J75)

Heat Exchanger (Cooling-Tower) Outlet (State 6). For the indirect cooling-

tower, the temperature drop across the cooling-tower is assumed ATc (see Table

6); therefore,
T6 = T5 — ATc
p6 =p5 (59)
he = h(Ts+P6)

Finally, COP for the enhanced cycle is obtained by

COP = ”—2 (60)
W0 +wpK
where
qL = “1 474
w,_ = h2 —h1 (51)
wp = h5 -h3

67

5.3. PERFORMANCE RESULTS AND DISCUSSIONS

Figure 33 shows the percentage increase in COP relative to baseline versus
the evaporator temperature (T1) for different mass flow ratios K. By increasing
evaporator temperature T1, the COP of the wave-rotor-enhanced cycle is
increased relative to the COP of the baseline cycle. This trend is seen until the
compressor pressure ratio in the enhanced cycle (l'lc=p2/p1) is reduced to a value
that is equal to the wave rotor pressure ratio ([76 =PRw). After that, the relative

COPga.n drops dramatically.

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

18 7 6 5 4n° s 245

.. $542.1: A
s / 3
$14 ////l
= . E
$12 $60 70/4 3
a 2". E
:10 V/P/l/+
TE 5//°°‘% // .ag
g 8 //, / g i

4 .

 

 

 

 

 

81012141618 20
T1W3]

Figure 33: Relative COP increase versus evaporation temperature for different
mass flow ratios

0
N
A
O)

68

Figure 34 represents the percentage increase in COP relative to baseline
versus mass flow ratio K for different evaporator temperatures, such as a side

view of Figure 33.

 

18

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

116° ”EFT?! .
16 , 1 " '
s: x’
’6' 15°C
.5 14
3
a
a
D /
8 12
3
+5 10°C
5 1O --------- .Ic'P/i"
8 / :
I o
.5 1 /____§_9
8 8 4’ .
g / 5 T, =o.1 c
0 l
.E 6 :
; PRw= 2.45
: T3= 326°C
4 I
o 100 200 300 400 500
I<=rir,,lm2
Figure 34: Relative COP increase versus mass flow ratio for different evaporation
temperatures

It shows only the increasing branches of Figure 33 up to an evaporation
temperature where the pressure ratio of the turbocompressor is reduced to the
value of the wave rotor pressure ratio. Increasing the mass flow ratio above 200
appears as ineffective according to Figure 34.

Figure 35 shows the effect of the wave rotor pressure ratio (PRw) on the
percentage increase in COP relative to baseline for different mass flow ratios K.

Each curve has a maximum point that indicates the best choice of wave rotor

69

pressure ratio for the given system Specifications. The location of this point
depends on several parameters including the hydraulic efficiency of the pump,
compressor pclytropic efficiency, evaporator temperature, and temperature

IiftT3 - T1, but not the mass flow ratio.

 

  

 

 
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

12
3 $00 13ch
g 10 fi\ ............. - 3 .
g 3 //,/3—'\\\§
z , 5 a)
(V N
a i .4 \
g 4 // ‘°\\k
2 2 \\
5 E \
o 2 2.5 3 3.5 4 4'5
PRw

Figure 35:Relative COP increase versus the wave rotor pressure ratio for
different mass flow ratios.

One common characteristic shown in Figure 33 to Figure 35 is that a
continued increase of the independent value does not always increase the
COPgaln. Although Figure 35 shows this effect for the wave rotor pressure ratio,
Figure 33 reveals a growing gradient of COPga.n up to the point where further
increase of evaporator temperature actually reduces COPgaln. A similar trend can
be seen in Figure 34 where the curves have an asymptotic behavior for
increased mass flow ratio.

Figure 36 shows the heat rejecter temperature T3 versus evaporator

temperature for different wave rotor pressure ratios and a constant relative

70

COPWn of 10%. The figure indicates that there are several options for the wave
rotor pressure ratio to obtain a certain relative COPgam. However, only the

optimum pressure ratio of 2.45 yields the highest temperature lift.

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

80
K=125
7O . increase ln COP relative
to baseline cycle =10% 1.0:" (2,5
x
2

60 ..

50 /,///A
_. {yé 4
g.” 40 I I/ V
'—

P /
so """" T " / /
//f / PRw=2.45
20 ¢/// 2 2.5
1 5 2.75
10 ./// 3
/ : 1.25
o ‘ l .
O 5 10 15 20 25 30 35 40 45

T1l°cl

Figure 36: Heat rejecter temperature versus evaporator temperature for different
wave rotor pressure ratios.

Figure 37 is a performance map of the enhanced cycle. Each point on this
plot shows the maximum COPE,“n that can be obtained by the optimum choice of
PRw for a given evaporator temperature and temperature lift. The constant PRw
lines indicate the optimum PRw that yields the highest possible COP”... indicated

by constant maximum COP“.n lines.

71

 
     
 

2519.90
4917 179.9 2840.90 9.14
4917 32.6 136.50 0.47
1227 10.0 136.50 0.49

 

 

 

 

 

 

 

 

Table 7: Calculated baseline cycle state values for the selected optimum point P

1227 10.0 2519.90 8.90
2007 61.9 2616.61 8.98
4917 154.0 2791.27 9.03

4917 32.6 143.77 0.49
6307 34.7 145.33 0.50
6307 31.7 123.99 0.43
1227 10.0 143.77 0.51

 

Table 8: Calculated enhanced-cycle state values for the selected optimum point P

In this performance plot of optimized points, an arbitrary optimum point P is
marked. This point is used to show the connection between all the performance
graphs. However, it is the only optimum point in the performance plots of Figure
33 to Figure 36. All other points of these plots are not found in

Figure 37 because they show a smaller COP”... than the points in

Figure 37. The calculated state values for the optimum point P are tabulated

in Table 7 and Table 8.

72

be seen in the performance map of the optimized points (Figure 37) by moving
point P to the right along constant PRw line of 2.45. However, for temperature
lifts below ~15 K (above maximum COPgain of ~16%) this effect is reversed such
that the maximum COPgain decreases with increasing evaporator temperature.
Figure 35 showed that for a given mass flow ratio and a combination of
evaporation and heat rejecter temperatures maximum COPgam exists. Since

point P is at the maximum of the dome, it also appears in Figure 37.

Max. increase in COP relatlve to baseline cycle

in,
I."
4'"
‘-

PRW=2'3

 

0 5 10 1 5 20 25 30 35 40 45
are]

Figure 37: Perforrnanoe map: maximum performance increase and optimum
wave rotor pressure ratios.

73

The following results can be obtained from the performance map of optimized
points:
1. The lower the temperature lift, the higher the relative maximum
COPgain of the enhanced cycle is.
2. The maximum optimum PRw for a condensing 3-port wave rotor is
about 2.51.
3.Although below a temperature lift of approximately 18 K, the optimum
PRw increases with the temperature lift and decreases slightly for higher
temperature lifts. This trend shows a rapidly decreasing gradient with
increasing temperature lift.
4. The slope of constant Copgain lines increases as the temperature lift
increases, showing that an increase of evaporation temperature is even more
beneficial for the enhanced cycle, especially for greater temperature lifts as it

is already the case for the R718 baseline cycle.

74

5. EXPERIMENTAL STUDY OF SHOCK WAVE VISUALIZATION USING
SCHLIEREN METHOD

5.1. SCHLIEREN VISUALIZATION TECHNIQUE

Schlieren methods visualize changes in the refractive index of transparent
fluids, and have applications in many areas of flow visualization especially shock
wave studies. The Schlieren method holds a special place in flow visualization
because it produces a natural, easily interpretable image of the refractive index
of gradient fields. A Schlieren instrument records changes in the refractive index
distribution of transparent media-like air flows. The refractive index distribution
can then be related to density, temperature, or pressure distributions within the
flow. The basic system uses light with parallel rays with a plane wave front. In
order to have parallel rays, the source is passed through a thin slit, and then
collimated by either a mirror or a lens. The collimated light is passed through the
flow and then brought to focus. If the flow is steady, the focal spot is an image of
the source slit; but if any turbulence exists, scintillation will occur. In order to
intensify the effect, a knife edge is placed in front of the camera, positioned to
block off about half of the light. In uniform flow this will simply make the
photograph half as bright. However in turbulent flow the light's path will be bent,
because of changes in the index of refraction caused by density gradients. In this
case some of the light that would othewvise be seen in the camera will instead hit
the knife-edge, and some that would normally be blocked will become visible.
The result is a set of lighter and darker patches corresponding to different

densities in the fluid. A simple Schlieren system can be built with a single mirror

75

with a long focal length; telescope mirrors can be used for this purpose. The
camera and light source are placed close together on either side of the focus,

and are aimed at the mirror (Figure 38).

 

 

 

 

 

parabolic
mirror
* half mirror .
_ - .3. ~‘= * '3'" light source
5: ~ ............ ‘— X
L Ki- : :‘ .tfi...+l .~‘:§. ........
.é :
1;, lens
‘0
a ff
test section ‘- cut 0
8
3
CD
3

Figure 38: Diagram of Schlieren setup for shock visualization

5.2. SCHLIEREN TECHNIQUE FOR VISUALIZATION OF SHOCK WAVE IN A SMALL SIZE

WAVE ROTOR

In order to understand the shock wave propagation in wave rotor channels
and examine the effectiveness of Shlieren technique for shock visualization in
wave rotor channels, experimental equipment was built at University of Tokyo to
study the fluid flow inside a single channel of a miniaturized wave rotor. For
visualization, the Schlieren technique was used to study the shock wave traveling
in a miniaturized channel of a wave rotor. Figure 39 shows the Schlieren setup at

University of Tokyo.

76

Light Source

_” _ ”was: ~
I..~~»’

Camera Half ’37]! ror

l I *, 1;:
Test Section
‘ \

 

Figure 39: Schlieren setup at University of Tokyo for shock wave visualization in
a single channel of a wave rotor

In this special design, the design of a stationary single channel is so that it
simulates the flow dynamics of a rotating channel. For this purpose, the buffer
chamber is fixed, and high pressure air coming from a screw compressor is
supplied constantly. To introduce the high pressure air to the test section
periodically, a rotor with inlet and outlet ports is rotating. A pressure transducer is
used to monitor the pressure change at the end of the tube (Figure 40). Also, in
order to trigger the camera, the passage openings were determined using a laser
which was pointing at the rotating shaft. The rotating shaft was painted so that it
was half black and white. Therefore, the reflection of the laser from these two

colored areas was different and could be detected.

77

 

Figure 40: Especial design for fluid study in a single channel of a wave rotor
(University of Tokyo)

To study the flow inside the miniaturized wave rotors, two glass-made cells
with the lengths of 38 mm and 165 mm with rectangular cross sectional area of
3mm by 3mm were used. Figure 41 shows a photograph of the 38 mm glass-cell

used for this study.

 

Figure 41: A 38 mm Glass-made channel for Shock Visualization

78

In the wave rotors 1.’ is defined as the ratio of passage opening time and the
wave travel time. To compare the pressure distribution at the end wall of the test
sections, cases with the same 1: value were considered. Rotor rotational speed
of 40 Hz for the short cross section is corresponding to the 10 Hz rotor rotational
speed for the long test section. As shown in Figure 42, the gradient of pressure
jump for both test sections up to pressure ratio of 0.25 MPa is the same. This
pressure ratio is corresponding to the primary shock wave. After this point the
short cross section has sharper gradient. This difference in the gradient could be
because of the size scale which results in different shock interactions such as
arrival of expansion wave. The fluctuations in the short cross section are mainly

because of the oscillatory nature of pressure injection by the compressor.

 

I

Pressure [MPa]
0 O
u b

P
we

......

 

0.1

 

 

 

 

-0.002 -0.001 0 0.001 0.002 0.003 0.004
Time [s]

Figure 42: Pressure distribution at the end wall of the short and long test sections
(17 =2.5, A= 3*3)

79

Figure 43 shows the picture obtained using the Schlieren method for the 38mm
test section. In this picture, the primary and secondary shock waves are clearly
observed. The speed of the primary shock is 470 m/s and the secondary shock
315 m/s which are slightly higher than the calculated values. This could be
because of the error of measurement devices such as the pressure gage. In
Figure 42, the start of pressure jump is at 470 microseconds. This fact can be

observed in Figure 43 as the primary shock wave reaches to the end and is

reflected at 470 microseconds.

U'IU'I
QV
0°
'E'C
rn (n

srl 099
srl 099
srl 079
srl 099
srl 029
9“ 0L9
srl 009
90 06v
srl 09v
srl on
srl 09v
9“ 097
Sn ow

..I
1"

38 mm

 

 

I
\

Figure 43: Shock wave visualization using Schlieren method for a 38 mm test
section (rr =2.5, N=40 Hz, A= 3*3)

Figure 44 is the same observation for the longer test section. During the
experiment, it was observed that the higher the rotational speed, the easier it was

to visualize the shock wave.

Also, for the longer test section, visualization of the primary shock using

Schlieren method was not possible, and as it can be seen in the picture only the

secondary shock is observable.

80

500 us
550 us
600 us
600 us
700 us
750 us
800 us
850 us

 

 

I‘ ,.

165 mm

Figure 44: Shock wave visualization using Schlieren Method for 165 mm Test
Section ((Tl' =2.5, N=40 Hz, A= 3*3)

Although the Schlieren technique was only used in the gaseous phase, the
same methodology is applicable in the condensing wave rotor. The major
difference is that in the condensing wave rotor the high pressure vapor is
produced by the means of flash evaporation. The experimental analysis of shock
wave visualization in the channels of a condensing wave rotor remains for future

work.

81

SUMMARY

In the present study, the advantages of cycles working with water as a
refrigerant (R718) and challenges involved with designing them are mentioned. A
code has been developed to compare R718 with other refrigerants in p-h, T-s,
and p-T diagrams. Also, the code generates the COP isolines for R718 and a
selected refrigerant for investigation of the best refrigerant choice in any
operating point.

To enhance the turbocompression and improve the efficiency of R718
cycles, the novel concept of 3-port condensing wave rotors integrated in R718
compression refrigeration cycles is investigated. The condensing wave rotor
employs pressurized water to pressurize, desuperheat, and condense the
superheated vapor coming from the compressor, all in one dynamic process. The
schematic p-h and T-s diagrams of the external processes and the wave and
phase-change diagram of the internal process are discussed. Flash evaporation,
shock wave compression, desuperheating, and condensation phenomena inside
the wave rotor channels are described. A computer code based on a
thermodynamic model is developed to evaluate the performance improvement of
R718 enhanced cycles. The effect of some key parameters on the performance
enhancement is demonstrated as an aid for optimization. Finally, a performance
map showing the optimized points of the enhanced cycle is presented. The
presented results show an additional improvement of the COP of up to 22% by
using the 3-port condensing wave rotor. Besides the performance enhancement,

the condensing wave rotor allows lower compressor pressure ratios for the same

82

temperature lift or increases the temperature lift without changing the
compressor.

This wave rotor is a simple drum, easy to manufacture, rotating at relatively
low speed. Because it performs compression, desuperheating, and condensation
in one compact device, it can reduce the size and cost of modern state-of-the-art
R718-chillers that now employ high-tech multistage compressors, intercooler,
and relatively bulky condensers.

In the last chapter, as a proof of concept, an experimental investigation by the
author for flow visualization in the channels of a wave rotor is described. The
shock wave propagation inside the channels of a small size wave rotor is
visualized using Shlieren technique. The same concept is suggested for future
work to visualize the shock propagation inside channels of a condensing wave

rotor.

83

REFERENCES

[1] Yunas, C., Boles, M, 1998,” Thermodynamics: An Engineering Approach”

[2]

[3]

[4]

[5]

[6]

[8]

[9]

[10]

[11]

[12]

[13]

Highstown: McGraw Hill.

Rowland, F. S., Molina, M. J., 1994 “Ozone Depletion: 20 Years after the
Alarm," Chemical and Engineering News, Vol. 72, pp: 8-13.

US. Environmental Protection Agency Web Site: Health Effects of
Overexposure to the Sun http://www.epa.gov/sunwiseluvandhealt.html.

Ozone Secretariat United Nations Environment Programme, 2001, The
Vienna Convention for the Protection of the Ozone Layer, UNON, Kenya.

Anderson, S. O., Sarma, K. M., 2002, Protecting the Ozone Layer,
Earthscan Publications Ltd, US.

Griffith, G., 1993, “Alternate Refrigerants for the Refrigeration Industry”
ASHREA-NIST Refrigerants Conference, pp.19-28.

Parson, E. A., 2003, Protecting the Ozone Layer: Science and Strategy,
Oxford University Press, New York.

Gerwen R. J. M., Verchoor M. J. E., 1997 “Air as Refrigerant: Prospects and
Challenges,” IEA Workshop Proceedings, Annex 22, Gatlinburg,
Tennessee, pp: 87-93.

Giles, G. R., Hunt, R. G., and Stevenson, G. F., 1997 “Air as a Refrigerant
for the 21st Century,” Proceedings of ASHRAE-NIST Conference,
Gaithersburg, US, pp: 96-103.

Vrinat, G., 2003 “Ammonia: Technical Characteristics, Regulations and
Impact on the Environment,” Chaud Froid Plomb.(French Journal), 658, pp:
55-60.

The US. Occupational Health and Safety Administration (OSHA) Laws:
1998 - 06/26/1998 - Storage and Handling of Anhydrous Ammonia (29 CFR
1910.111), www.osha.gov.

Granryd, E., 2001 “Hydrocarbons as Refrigerants: an Overview,”
lntemational Journal of Refrigeration, 24, No. 1, pp: 15-24.

Pearson, S. F., 1996 “Uses of Hydrocarbon Refrigerants,” lntemational
Institute of Refrigeration, Proceeding of Aarhus Meeting, pp: 439-446.

84

[14] Orbitz, T. M., Li, D., Groll E. A., 2003 “Evaluation of the Performance of
Potential of COZ as a Refrigerant in Air-to-Air Air Conditioners and Heat
Pumps: System Modeling and Analysis,” ARTI-21CR/610-10030.

[15] Hans, J. H., Hwang, Y., and Raderrnacher R., 2002 “Options for A Two-
Stage Transcritical Carbon Dioxide Cycle,” Proceeding of IIR conference
Commissions B1, 82, E1 and E2, Guangzhou, China, pp: 158-164.

[16] Miller N., 2002, “Turbo Chillers using Water as a Refrigerant” ASME
Process Industry Division PID Newsletter, Fall 2002, p 3.

[17] Kharazi, A. A., Akbari, P., and Muller, N., 2004, “An Application of Wave
Rotor Technology for Performance Enhancement of R718 Refrigeration
Cycles” 2nd lntemational Energy Conversion Engineering Conference, AIAA
Paper 2004-5636.

[18] Kharazi, A. A., Akbari, P., and Mliller, N., 2004, “Performance Benefits of
R718 Turbo-Compression Cycles Using a 3-Port Condensing Wave Rotors”
2004 lntemational Mechanical Engineering Conference, ASME Paper
IMECE2004- 609926.

[19] Yuan, Q.S., Blaise, J.C., 1998, ”Water - A Working Fluid for CFC
Replacement,” Rev. Int. Froid, 11, pp. 243-248.

[20] Kilicarslan, A., Miller, N., 2004, “COPs of R718 in Comparison with Other
Modern Refrigerants” The First Cappadocia lntemational Mechanical
Engineering Symposium, Cappadocia, Turkey.

[21] Rossi, F., Mastrullo, R 1991, “Working Fluids Thermodynamic Behavior for
Vapor Compression Cycles" Applied Energy 0306-2619, pp.163-180.

[22] Albring, P., 1994, “Water as A Refrigerant in Refrigeration Plants with
Mechanical Compression” ISSN 0151-1637, New Applications of the Natural
Working Fluids in Refrigeration and Air Conditioning, IIR, Hannover, pp.
735-742.

[23] Albring, P., Heinrich, G., 1998, “Turbo Chiller with Water as Refrigerant”
Natural Working Fluids ’98, IIR, Oslo, pp. 93-103.

[24] Lachner, B., Nellis, G., and Reindl, D., 2004, “An Investigation into the
Feasibility of the Use of Water as a Refrigerant” lntemational Refrigeration
and Air Conditioning Conference, Purdue, R160, pp. 1-8.

[25] Miller, N., 2001, “Design of Compressor Impellers for Water as a
Refrigerant” ASHRAE Transactions, 107, pp. 214-222.

85

[26] Wobst, E., Kalitzin, N., and Apley, R., 2005, “Turbo Water Chiller with Water
as Refrigerant” Seventeenth lntemational Compressor Engineering
Conference at Purdue, C129.

[27] ILK Dresden, Germany, Company Brochure, 2004, “R718-Turbo-Chiller
System Dimensions" R718-Turbo-A626 (ATV110/1-25-52 and ATV110/3-
21-52).

[28] Gupta, R. P., Linton, J. W., and Snelson, W. K., 1994 “R&D Test facility to
Develop Systems Using Water as Refrigerant,”, IEA, Annex 22 / IEA HPP,
Trondheim Workshop Proc., pp: 203-215.

[29] Wobst, E., Kalitzin, N., Apley, R., 2004 “Turbo Water Chiller with Water as
Refrigerant”, lntemational Compressor Engineering Conference at Purdue,
Paper C129.

[30] Paul, J., Jahn E., 1996, “Water Chilling and Ice Making with Water as
Refrigerant,” Preceedings of the lntemational Conference on Ozon
Protection Technologies, Washington, USA, pp: 955-960.

[31] Nyvad J., Elefsen F., 1993, “Energy Efficient Cooling by Use of Cycloid
Water Vapour Compressor,” IIR, Proceedings of Ghent Meeting, pp: 67-74.

[32] Hackensellner T., Jurisch C., 1994, “Water: A Working Fluid for a
Compression Heat Pump with a Liquid Ring Compressor,” IIR, Proceedings
of Hanover Meeting, pp: 755-764.

[33] Grazzini, G., D'Albero, M. 1998, “A Jet-Pump Inverse Cycle with Water
Pumping Column," Proceeding of Meeting of Commissions B1, 32, E1, E2,
IIR, Oslo, pp. 63—68.

[34] Wight, S. E., Yoshinaka, T., Le Drew, B. A., and D’Orsi, N. C., 2000 “T he
Efficiency Limits of Water Vapor Compressors,” ARTl-21CR/605-10010-01.

[35] Paul, J., 1998 “District Cooling with Water as Refrigerant Combined with
Energy Storage Systems,” IIF-IIR, Commission B2,with B1,E1&E2, Oslo,
Norway, pp.82-92.

[36] Zimerrnan, 2., 1994 “Development of Large Capacity High Efficiency
Mechanical Compression (MVC) Units,” Desalination, 96, pp: 51-58.

[37] Madsboll, M. S., Minds, 6., Nyvad, J., and Elefsen, F., 1994 “The State of

Art for Water Vapour Compressors and Cooling Plants Using Water as
Refrigerant,” Proc. Hanover Meet, IIR, pp: 743-758.

86

[38] Albring, P., Burandt B.,and Schenk, N., 2002, “Water Vapour Compression:

[39]

Application and Outlook,” KI Luft Kaltetech DE, 38, n.6, pp: 276-281.

KiJhnI-Kinel, J., 1997 “New Age Water Chillers with Water as Refrigerant,”
CERN, Geneva.

[40] Weber, H. E., 1995, Shock Wave Engine Design, John Wiley and Sons,

New York.

[41] Weber, H. E., 1986, “Shock-Expansion Wave Engines: New Directions for

Power Production” ASME Paper 86-GT-62.

[42] Akbari, P., Kharazi, A. A., and Miiller, N., 2003, “Utilizing Wave Rotor

[43]

[44]

[45]

[46]

[47]

[48]

[49]

[50]

Technology to Enhance the Turbo Compression in Power and Refrigeration
Cycles” 2003 lntemational Mechanical Engineering Conference, ASME
Paper lMECE2003-44222.

Akbari, P.,Nalim, M. R., and Muller, N., 2004, “A Review of Wave Rotor
Technology and Recent Developments” 2004 lntemational Mechanical
Engineering Conference, ASME Paper lMECE2004-60082, USA.

Taussig, R. T., Hertzberg, A., 1984, “Wave Rotors for Turbomachinery“
Winter Annual Meeting of the ASME, edited by Sladky, J. F., Machinery for
Direct Fluid-Fluid Energy Exchange, AD-07, pp. 1-7.

Shreeve, R. P., Mathur, A., 1985, Proceeding ONRINAVAIR Wave Rotor
Research and Technology Workshop, Report NPS-67-85-008, Naval
Postgraduate School, Monterey, CA.

Paxson, D. E., 1995, “Comparison Between Numerically Modeled and
Experimentally Measured Wave-Rotor Loss Mechanism” Journal of
Propulsion and Power, 11, No. 5, pp. 908-914. '

Paxson, D. E., 1996, “Numerical Simulation of Dynamic Wave Rotor
Performance” Journal of Propulsion and Power, 12, No. 5, pp. 949-957. See
also AIAA Paper 95-2800 and NASA TM 106997.

Wilson, J., Paxson, D. E., 1996, “Wave Rotor Optimization for Gas Turbine
Topping Cycles” Journal of Propulsion and Power, 12, No.4, pp. 778-785.

Welch, G. E., Jones, S. M., and Paxson, D. E., 1997, “Wave Rotor-
Enhanced Gas Turbine Engines” Journal of Engineering for Gas Turbines
and Power, 1 19, No. 2, pp. 469-477.

Welch, G. E., 1997, “Macroscopic Balance Model for Wave Rotors” Journal
of Propulsion and Power, 13, No. 4, pp. 508-516.

87

[51] Welch, G. E., 1997, “Two-Dimensional Computational Model for Wave Rotor
Flow Dynamics” Journal Engineering for Gas Turbines and Power, 119, No.
4. pp. 978-985. See also ASMA Paper 96-GT-550 and NASA TM 107192.

[52] Vlfilson, J., 1998, “An Experimental Determination 0f Loses in a Three-Port
Wave Rotor” Journal of Engineering for Gas Turbines and Power, 120, pp.
833-842.

[53] Paxson, D. E., Nalim, M. R., 1999, “Modified Through-Flow Wave-Rotor
Cycle with Combustor Bypass Ducts” Journal of Propulsion and Power, 15,
No. 3, pp. 462-467.

[54] Welch, G. E., 2000, “Overview of Wave-Rotor Technology for Gas Turbine
Engine Topping Cycles” Novel Aero Propulsion Systems lntemational
Symposium, The Institution of Mechanical Engineers, pp. 2-17.

[55] Smith, C. F., Snyder, P. H., Emmerson, C. W., Nalim M. R., 2002, “Impact
of the Constant Volume Combustor on a Supersonic Turbofan Engine" AIAA
Paper 2002-3916.

[56] Akbari, P., and Muller, N., 2005, “Wave Rotor Research Program at
Michigan State University” 41St AIAA/ASME/SAE/ASEE Joint Propulsion
Conference and Exhibit, AIAA Paper 2005—3844.

[57] Saury, D., Harrnand, S., and Siroux, M, 2001, “Experimental Study of Flash
Evaporation of Water F ilm” lntemational Journal of Heat and Mass Transfer,
No.45, pp. 3447-3457.

[58] Kentfield, J. A. C., 1998, “Wave Rotors and Highlights of Their
Development,” AIAA Paper 98-3248.

[59] Burghard, H., 1929, German Patent 485386.
[60] Darrieus, G., 1950, “Pressure Exchange Apparatus,” US. Patent 2526618.

[61] Kentfield, J. A. C., 1993, Nonsteady, One-Dimensional, Internal,
Compressible Flows, Oxford University Press, Oxford.

[62] Gyarrnathy, G., 1983, “How Does the Comprex® Pressure-Wave
Supercharger Work?” SAE Paper 830234.

[63] Zehnder, G., Mayer, A. and Mathews, L., 1989, “The Free Running
Comprex®,” SAE Paper 890452

88

[64] Kharazi, A. A., Akbari, P., and Muller, N., 2004, “Preliminary Study of a
Novel R718 Turbo-Compression Cycle Using a 3-Port Condensing Wave
Rotor” 2004 lntemational ASME Turbo Exposition, ASME Paper GT2004-
53622.

[65] Miyatake, O., Murakami, K., and Kawata, Y., 1973, “Fundamental
Experiments with Flash Evaporation” Heat Transfer Jpn., Res. 2, pp. 89-
100.

[66] Anderson, J. D., 2003, Modern Compressible Flow, McGraw hill, Third
Edition.

[67] Daneshyar, H., 1976, “One-Dimensional Compressible Flow" Pergamon
Press.

[68] Griffith, G., 1993 “Alternate Refrigerants for the Refrigeration Industry,”
ASHREA-NIST Refrigerants Conference. 09.19-28.

[69] Parson, E. A., 2003, Protecting the Ozone Layer: Science and Strategy,
Oxford University Press, New York, US.

[70] Albring, P., Heinrich, G., 1996, “R718 Heat Pumps: Applications for Natural
Refrigerants,” IIR, Aarhus, pp. 553-558.

[71] Miller, N., 2001 “Design of compressor impellers for water as a refrigerant”
ASHRAE Transaction 107, pp: 214-222.

89

LIBR

i[iilijijiiijjjiiiiiiiji