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Experimental Study of a Novel Gas Turbine Engine Integrated with an Absorption Refrigeration System

Permanent Link: http://ufdc.ufl.edu/UFE0021324/00001

Material Information

Title: Experimental Study of a Novel Gas Turbine Engine Integrated with an Absorption Refrigeration System
Physical Description: 1 online resource (137 p.)
Language: english
Creator: Howell, Eric Baker
Publisher: University of Florida
Place of Publication: Gainesville, Fla.
Publication Date: 2007

Subjects

Subjects / Keywords: Mechanical and Aerospace Engineering -- Dissertations, Academic -- UF
Genre: Mechanical Engineering thesis, M.S.
bibliography   ( marcgt )
theses   ( marcgt )
government publication (state, provincial, terriorial, dependent)   ( marcgt )
born-digital   ( sobekcm )
Electronic Thesis or Dissertation

Notes

Abstract: A series of tests were performed on a novel gas turbine engine, coupled to a vapor absorption refrigeration system. The engine cycle is unique in that it is semi-closed and turbocharged in addition to utilizing a recuperator. The refrigeration system is in place to remove heat from the recirculating exhaust gases, to cool the engine inlet gases, and to produce additional refrigeration for auxiliary thermal loads. The engine was developed in several stages and operated over a wide range of conditions. The data from these experiments were reduced, analyzed and presented with emphasis placed on the unique attributes of the combined cycle. Observed hardware limitations are also identified, and adjustments were made to the data in an effort to estimate certain performance parameters for more ideal components. It was observed in these experiments that power transitions could be made with this cycle by varying the turbocharger pressure ratio. Doing so results in a somewhat constant thermal efficiency over a range of power levels, since the core engine approximately remains at a constant, dimensionless operating point. It is also shown that semi-closure of the power cycle reduces the air flow necessary to run the engine, highlighting the anticipated size reduction of certain components. The data also validate that semi-closure of the cycle serves to alter the combustion environment, such that NOX emissions are reduced, making the power cycle more attractive from an environmental perspective. The cooling of recirculation gases presents the additional opportunity to extract fresh water from the gas path. This capability is also realized in these experiments, but the amount of water extracted was limited by some of the experimental hardware to less than half the theoretical maximum.
General Note: In the series University of Florida Digital Collections.
General Note: Includes vita.
Bibliography: Includes bibliographical references.
Source of Description: Description based on online resource; title from PDF title page.
Source of Description: This bibliographic record is available under the Creative Commons CC0 public domain dedication. The University of Florida Libraries, as creator of this bibliographic record, has waived all rights to it worldwide under copyright law, including all related and neighboring rights, to the extent allowed by law.
Statement of Responsibility: by Eric Baker Howell.
Thesis: Thesis (M.S.)--University of Florida, 2007.
Local: Adviser: Lear, William E.

Record Information

Source Institution: UFRGP
Rights Management: Applicable rights reserved.
Classification: lcc - LD1780 2007
System ID: UFE0021324:00001

Permanent Link: http://ufdc.ufl.edu/UFE0021324/00001

Material Information

Title: Experimental Study of a Novel Gas Turbine Engine Integrated with an Absorption Refrigeration System
Physical Description: 1 online resource (137 p.)
Language: english
Creator: Howell, Eric Baker
Publisher: University of Florida
Place of Publication: Gainesville, Fla.
Publication Date: 2007

Subjects

Subjects / Keywords: Mechanical and Aerospace Engineering -- Dissertations, Academic -- UF
Genre: Mechanical Engineering thesis, M.S.
bibliography   ( marcgt )
theses   ( marcgt )
government publication (state, provincial, terriorial, dependent)   ( marcgt )
born-digital   ( sobekcm )
Electronic Thesis or Dissertation

Notes

Abstract: A series of tests were performed on a novel gas turbine engine, coupled to a vapor absorption refrigeration system. The engine cycle is unique in that it is semi-closed and turbocharged in addition to utilizing a recuperator. The refrigeration system is in place to remove heat from the recirculating exhaust gases, to cool the engine inlet gases, and to produce additional refrigeration for auxiliary thermal loads. The engine was developed in several stages and operated over a wide range of conditions. The data from these experiments were reduced, analyzed and presented with emphasis placed on the unique attributes of the combined cycle. Observed hardware limitations are also identified, and adjustments were made to the data in an effort to estimate certain performance parameters for more ideal components. It was observed in these experiments that power transitions could be made with this cycle by varying the turbocharger pressure ratio. Doing so results in a somewhat constant thermal efficiency over a range of power levels, since the core engine approximately remains at a constant, dimensionless operating point. It is also shown that semi-closure of the power cycle reduces the air flow necessary to run the engine, highlighting the anticipated size reduction of certain components. The data also validate that semi-closure of the cycle serves to alter the combustion environment, such that NOX emissions are reduced, making the power cycle more attractive from an environmental perspective. The cooling of recirculation gases presents the additional opportunity to extract fresh water from the gas path. This capability is also realized in these experiments, but the amount of water extracted was limited by some of the experimental hardware to less than half the theoretical maximum.
General Note: In the series University of Florida Digital Collections.
General Note: Includes vita.
Bibliography: Includes bibliographical references.
Source of Description: Description based on online resource; title from PDF title page.
Source of Description: This bibliographic record is available under the Creative Commons CC0 public domain dedication. The University of Florida Libraries, as creator of this bibliographic record, has waived all rights to it worldwide under copyright law, including all related and neighboring rights, to the extent allowed by law.
Statement of Responsibility: by Eric Baker Howell.
Thesis: Thesis (M.S.)--University of Florida, 2007.
Local: Adviser: Lear, William E.

Record Information

Source Institution: UFRGP
Rights Management: Applicable rights reserved.
Classification: lcc - LD1780 2007
System ID: UFE0021324:00001


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03067c9c3e7a5574cb7c29d83d9bc843
e314093b426c8dd0c5b83050748ddbcd55483a13







EXPERIMENTAL STUDY OF A NOVEL GAS TURBINE ENGINE INTEGRATED WITH
AN AB SORPTION REFRIGERATION SYSTEM





















by

ERIC B. HOWELL


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

UNIVERSITY OF FLORIDA

2007




































0 2007

Eric B. Howell










ACKNOWLEDGMENTS

I first thank my committee chair, Dr. William Lear, for giving me the great opportunity to

be a part of this research effort, and for his counsel throughout my graduate studies. I would also

thank Dr. S. A. Sherif, and Dr. Herbert Ingley for serving on my committee and demonstrating

great patience with me throughout this work. I must also thank Mr. John Crittenden for all of his

advice, encouragement, and comradeship during this undertaking. I am also indebted to Mr. Dan

Brown and Mr. Todd Nemec for all their technical assistance and recommendations.












TABLE OF CONTENTS


page

ACKNOWLEDGMENTS .............. ...............3.....


LIST OF TABLES ............ ..... ._ ...............7....


LIST OF FIGURES .............. ...............8.....


LIST OF TERMS ............ ..... ..__ ...............10...


AB S TRAC T ............._. .......... ..............._ 14...


CHAPTER



1 INTRODUCTION ................. ...............16.......... ......


2 LITERATURE REVIEW ................. ...............21................


History of Semi-Closed Cycle Concepts .............. ...............21....
Analytical Studies ................. ...............22.................
Experimental Studies ................ ...............24.................

3 TESTING HARDWARE ................. ...............26........... ....


Combined Cycle Components ................ ............ ...............26......
Engine Characteristics and Specifications............... .............2
Recuperator ................. ...............27.................
Turbocharger .............. ...............27....
Heat Exchangers ................. ...............27.................
Hot gas cooler (HGC) .............. ...............28....
Warm gas cooler (WGC)............... ...............28.
Cold gas cooler (CGC) ................. ...............28................
Dynamometer ................. ...............29.............
D ucting .............. ...............29....
Facility Resources............... ...............2
Chilled W ater. ............ ............ ...............29...
Fans and HVAC .............. ........ ............... .......3

Vapor Absorption Refrigeration System (VARS) ................. ...............30........... ...
Overview of Build Configurations ................. ...............30.......... ....
B uild 1 .............. ...............3 1....
Build 2A .............. ...............3 2....
B uild 3 .............. ...............33....
Build 4 .............. ...............3 4....












Data Measurement ............. ...... ._ ...............34...
Data Acquisition Hardware ............. ...... ...............34...
Data Acquisition Software .............. ...............36....
In strmentati on............_.._ ....._. ...............36...

Thermocouples ..........._.._ ....._. ...............36....
Pressure transducers .............. ...............37..

Turbine and paddle-wheel flowmeters .....__.....___ ..........._ ............3
Optical tachometer .............. ...............37....
Load cells .............. ...............38....
Gas analysis................ ...... .........3
Mechanical instruments and analog readers............... ...............39


4 TEST PROCEDURES ............. ...... .__ ...............45..


Test Plan ............... ...... .... .............4

Engine/Component Preparation ............. ...... .__ ...............45..
Instrument Preparation............... ..............4
Transducer Offsets............... ...............46
Ambient Conditions............... ...............4


5 DATA REDUCTION SCHEME ............. ...... .__ ...............49..


Gas Properties ............... ........_ ...............49....
Enthalpy and Specific Heat ................ ...............50................
Gas Composition .............. ...............52....
Basic Equations and Assumptions ................. ...............54................
Main Air Inlet (MAI)................... ...............5
Low Pressure Compressor (LPC) ................. ...............55................
Recirculation Venturi (RCV) .............. ...............55....
Hot Gas Cooler (HGC) ........._.__....... .__. ...............56...
Warm Gas Cooler (WGC) ........._.__........_. ...............56...
Cold Gas Cooler (CGC) .............. ...............58....
Ducting Section 4-5 ........._._............___ ...............59....
High Pressure Compressor (HPC) .......................__ ...............60. ...
Com bustor .............. ...............61....
Ducting Section 7-8 ................. ...............63....... ......

High Pressure Turbine (HPT) ................. ...............64....... .....
Recuperator. .............. .... ......._ ...............66.......
Low Pressure Turbine (LPT) ...._.. ................. ........_.. ........6
M ixing Junctions .............. ...............68....

Propagation of Uncertainty ................. ...............70...............
Data Adjustments............... ..............7
Pressure Drop Considerations............... .............7


6 EXPERIMENTAL RESULTS .............. ...............79....












7 CONCLUSIONS AND RECO1V1VENDATIONS ................ ...............................105


Summary ........._.___..... ._ __ ...............105....
Recommendations............... ............10


APPENDIX


A OPERATING PROCEDURES FOR HPRTE/VARS ............_...... .................110


B SETUP PROCEDURES FOR HPRTE/VARS ....__ ......_____ .......__ ...........13


C ADJUSTED EXPERI1VENTAL DATA ............__......___....._ ...........19


LIST OF REFERENCES ............ ..... ..__ ...............135..


BIOGRAPHICAL SKETCH ............ _...... ._ ...............137...












LIST OF TABLES


Table page


3-1. Temperature locations and instrumentation. ............. ...............40.....


3-2. Pressure locations and instrumentation. ............. ...............41.....


5-1. Uncertainty of instruments .............. ...............74....


5-2. Data scaling parameters [19]. ............. ...............75.....


C-1. Data from runs B4-1 and B4-2. ........._.._ ..... ._._ ...............119


C-2. Data from runs B4-3 and B4-5............... ...............121.


C-3. Data from runs B4-7 and B4-8.. ....__ ...._._ ......__.__......._._ ........._.....123


C-4. Data from runs B4-10 and B4-11 ..........._...__........ ...............125.


C-5. Data from run B4-12. ..........._..._ ...............127......._ ...


C-6. Data from run B4-14. ..........._..._ ...............129..._.._ ....


C-7. Data from runs B4-15 and B4-16. ..........._..._ ...............131...._.._ .


C-8. Data from run B4-17. ..........._.._ .................133._.._ ..











LIST OF FIGURES


Figure page

1-1. Block diagram of HPRTE flow path. ............. ...............20.....

3-1. Block diagram of Build 1 configuration. .............. ...............42....

3-2. Photograph of Build 1 improvised evaporator installation. ........... ......_. ..............43

3-3. Block diagram of Build 2A configuration. .............. ...............43....

3-4. Block diagram of Build 3 configuration ................. ...............44........... ..

3-5. Block diagram of Build 4 configuration. .........._._ ...._.._ .. ...............44..

5-1. Map used for Einding the HPC adiabatic exit temperature, shown as a function of
corrected flow rate and corrected speed [16]. ............. ...............76.....

5-2. Rover 1S-60 internal power losses as a function of rotor speed [16] ................. ...............77

6-1. Actual net power output from Build 4 engine runs. ............. ...............92.....

6-2. Corrected Power, power, and thermal efficiency over corrected speed. ............. ................93

6-3. Scaling parameters versus corrected speed. ............. ...............93.....

6-4. Theta versus LPC pressure ratio. ........._.._.. ...._... ...............94..

6-5. Power versus LPC pressure ratio. ........._.._.. ...._... ...............94..

6-6. High pressure turbomachinery temperature differences and mass flow rate versus LPC
pressure ratio. .............. ...............95....

6-7. Percent change in high pressure turbomachinery temperature differences and mass flow
rate versus LPC pressure ratio. ............. ...............95.....

6-8. Flow parameters versus corrected speed. ............. ...............96.....

6-9. Generalized effects of HPC pressure ratio and CTR on specific power, illustrated on a
T-S diagram ............. ...............96.....

6-10. Specific power versus CTR, for various HPC pressure ratios............... .................9

6-11. Corrected power versus CTR for various HPC pressure ratios .................... ...............9

6-12. HPC pressure ratio and CTR versus corrected speed. .............. ...............98....

6-13. Thermal efficiency versus CTR. ........... ......__ ...............98











6-14. Thermal efficiency versus HPC pressure ratio. .............. ...............99....

6-15. HPC isentropic efficiency versus corrected speed. ............. ...............99.....

6-16. Recirculation ratio versus LPC pressure ratio. ............. ...............100....

6-17. Fresh air flow rate versus recirculation ratio. .............. ...............100....

6-18. Equivalence ratio versus recirculation ratio. ............. ...............101....

6-19. Equivalence ratio versus LPC pressure ratio. .....__.....___ .......... .. .........0

6-20. Nitric oxide and carbon monoxide concentrations versus recirculation ratio. ..................102

6-21. Oxygen and nitrogen concentrations versus recirculation ratio. ................... ...............10

6-22. Water-to-fuel ratio versus LPC pressure ratio. ....._ .....___ ............_.........0

6-23. Recirculation flow rate versus LPC pressure ratio. ......___ .......__ ................103

6-24. Temperatures related to water extraction versus LPC pressure ratio ............... ...............104










LIST OF TERMS

Roman Letters

a first degree coefficient for specific heat polynomial

A cross sectional area

b second degree coefficient for specific heat polynomial

c third degree coefficient for specific heat polynomial

cP COnstant-pressure specific heat

e heat exchanger effectiveness

F force

FA fuel-to-air ratio

h specific enthalpy or convection coefficient

M molecular weight

thz mass flow rate

n moles

A mole flow rate

N engine speed

N* corrected engine speed

P pressure or power

P* corrected power

PR pressure ratio

Q heat transfer rate

R mass-specific gas constant or recirculation ratio

R mole-specific (universal) gas constant









T temperature

F volumetric flow rate

W power

y mole fraction

Greek Letters

Y specific heat ratio

6 pressure correction parameter

a emissivity

rl efficiency

6 temperature correction parameter

p density

a Stefan-Boltzmann constant

z dimensionless temperature parameter

to angular velocity, or uncertainty

Subscripts

0 reference state, 300 Kelvin

AMB ambient

AVG average

COLD cold side of heat exchanger

COMB combustor

CONV convective

DAQ data acquisition

DRY dry gas mixture, omitting water










DYNO dynamometer

fg liquid-vapor transition

HOT hot side of heat exchanger

HPTI high pressure turbine inlet

HPTX high pressure turbine exit

i iti' term in a series, or ith constituent of a mixture

in general inlet state point

INST instrumentation

LPCI low pressure compressor inlet

LPCX low pressure compressor exit

max maximum possible value

MIX mixture

MECH mechanical friction

NC non-condensable

out general outlet state point

R recuperator

RAD radiative

RCVI recirculation venturi inlet

SEN sensible

SURF surface

TOT total value for mixture

TRAN transient

W water










WGCG warm gas cooler, gas side

WGCI warm gas cooler inlet

WGCW warm gas cooler, water side

X generalized independent parameter









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

EXPERIMENTAL STUDY OF A NOVEL GAS TURBINE
ENGINE INTEGRATED WITH AN AB SORPTION REFRIGERATION SYSTEM



Eric B. Howell

August 2007

Chair: William E. Lear
Major: Mechanical Engineering

A series of tests were performed on a novel gas turbine engine, coupled to a vapor

absorption refrigeration system. The engine cycle is unique in that it is semi-closed and

turbocharged in addition to utilizing a recuperator. The refrigeration system is in place to

remove heat from the recirculating exhaust gases, to cool the engine inlet gases, and to produce

additional refrigeration for auxiliary thermal loads. The engine was developed in several stages

and operated over a wide range of conditions. The data from these experiments were reduced,

analyzed and presented with emphasis placed on the unique attributes of the combined cycle.

Observed hardware limitations are also identified, and adjustments were made to the data in an

effort to estimate certain performance parameters for more ideal components.

It was observed in these experiments that power transitions could be made with this cycle

by varying the turbocharger pressure ratio. Doing so results in a somewhat constant thermal

efficiency over a range of power levels, since the core engine approximately remains at a

constant, dimensionless operating point. It is also shown that semi-closure of the power cycle

reduces the air flow necessary to run the engine, highlighting the anticipated size reduction of

certain components. The data also validate that semi-closure of the cycle serves to alter the

combustion environment, such that NOx emissions are reduced, making the power cycle more










attractive from an environmental perspective. The cooling of recirculation gases presents the

additional opportunity to extract fresh water from the gas path. This capability is also realized in

these experiments, but the amount of water extracted was limited by some of the experimental

hardware to less than half the theoretical maximum.









CHAPTER 1
INTTRODUCTION

With rising energy costs and instability of energy sources, it is becoming ever more

important to develop newer technologies to meet the unrelenting increase in demand for energy.

This challenge is further complicated by concerns about the current and future states of the

environment, and must be overcome while concurrently decreasing any negative, environmental

impacts. Distributed generation of energy is seen as a potential solution, at least in part, to

present energy challenges. Some varieties of distributed generation exist in the form of solar

panels and wind turbines, which use natural energy sources that are often inconsistent and

dependant upon weather. Moreover, these types of systems must be implemented on a large

scale if they are to be significant contributors to the energy supply. Microturbines are also

becoming more attractive as distributed generation systems, as they are more compact and can

burn a large variety of fuels. Microturbines. also provide the capability of supporting heating

loads with exhaust gases, and more recent innovations utilize absorption refrigeration systems to

provide refrigeration loads as well.

It is hypothesized that several innovations to the conventional gas turbine engine can

increase its usefulness as a distributed generation system. The particular improvements

implemented in this study are believed to reduce the physical size of the engine and enable it to

follow load demands without significant sacrifices of thermal efficiency. In addition, it is

theoretically possible for the integration of such an engine with a Vapor Absorption

Refrigeration System (VARS) to further enhance its attractiveness by enabling the combined

cycle to produce external refrigeration, and fresh water.

In an effort to better quantify the extent of these prospective capabilities, modeling and

experimental activities were conducted at the University of Florida. This thesis details the










experimental efforts undertaken on a novel gas turbine engine cycle combined with a VARS.

The power cycle is unique in that it is both semi-closed and turbocharged, in addition to being

recuperated, and is referred to as the High Pressure Regenerative Turbine Engine (HPRTE).

Figure 1-1 shows the flow path of the HPRTE.

There are two main reasons for turbocharging the gas turbine engine. The first, more

apparent of these is the resulting increase in power density. By increasing the pressure of the

working fluid, the density is also increased. In turn, the same size engine can operate with a

greatly increased mass flow rate which equates to more power from a smaller engine. The

second reason for turbocharging the gas turbine engine is more subtle. The engine thermal

efficiency can remain approximately constant over a range of power levels. This is

accomplished by using the turbocharger itself to perform power transitions. Exhaust can be

diverted through/around the Low Pressure Turbine (LPT) to raise/lower the inlet pressure of the

engine. If the inlet temperature is held constant, then the density and flow rate will in turn

increase proportionately to the inlet pressure. Doing so allows the core engine to remain at the

same, dimensionless operating point (same corrected flow corrected power and pressure ratios)

for different power levels. The core engine is simply the recuperated gas turbine engine being

operated apart from other, supporting hardware (turbocharger, VARS, etc.).

It is also theorized that the semi-closure of the gas turbine cycle also offers several

advantages over conventional gas turbine cycles. The first of these relates to the fact that about

20 percent of the air ingested by a conventional gas turbine engine is necessary to support

combustion. Semi-closing the cycle entails recirculating much of the exhaust gases to be used

again as the working fluid, and as this recirculated proportion of gas is increased, the net inflow

and outflow of air and exhaust gas are reduced. This effect is believed to significantly reduce the









size of the turbocharger and other inlet air and exhaust ducting. Some applications may also Eind

significant utility in the smaller thermal and acoustic signatures from the engine resulting from

the reduced exhaust flow.

Speculations are also made regarding the attractiveness of the semi-closed gas turbine

engine from an environmental perspective. The more oxygen-dilute gases are expected to reduce

the temperature within the primary zone of the combustor. This should result in a significant

reduction of oxides of nitrogen (NOx), since NOx formation rates scale very strongly with

temperature. Similar arguments can also be made about soot formation rates.

Recirculating significant quantities of hot exhaust gases to the High Pressure Compressor

(HPC) inlet presents both the problem of significant cooling, and the benefit of fresh water

extraction. Aside from its front-end material considerations, the gas turbine engine would

perform poorly with such extreme temperatures of mixed inlet gases. Hence, these gases must

be cooled. However, a benefit is made available by the cooling of these gases, as an appreciable

fraction of the recirculation flow is water vapor. Once condensed in the cooling process, the

water can straightforwardly be extracted and made ready to drink. This rather large cooling

challenge presented by the recirculating exhaust gases can be viewed as an opportunity to extract

further potential from the power cycle via integration of the VARS.

The VARS is driven by the heat rej ected from the recirculation loop and in turn benefits

the HPRTE in several ways. The most notable of these is the theoretical capability it offers for

reducing the HPC inlet temperature to below ambient temperatures, and improving the engine

performance. This attribute implies that the engine performance should depend much less, if at

all, on ambient conditions, since the HPC inlet temperature and pressure can be controlled by the

VARS and turbocharger, respectively. Additionally, it is theorized that the VARS can produce









more refrigeration than is necessary for the gas-path cooling. As a result, a fourth commodity,

refrigeration, can be counted alongside of power, heat, and water as an output from this

combined cycle.

The apparent advantages and capabilities of this combined cycle motivated the US Army

Research Laboratory to fund the development of an experimental HPRTE/VARS combined

cycle for the purpose of demonstrating its thermodynamic principles. The University of Florida

was subcontracted to carry out the design and testing of several experimental apparatuses within

its Energy and Gasdynamics Laboratory (EGDL). The overall objectives of this research work

were to assemble an experimental HPRTE/VARS combined cycle, demonstrate its operability,

and demonstrate that its theoretical capabilities could be realized in accordance with the

predictions of thermodynamic models. A stepwise approach was taken in the experimental

development such that the capabilities of the experimental apparatus were expanded

incrementally. A second objective of the experimental work was to help develop and validate

design-point models for this combined cycle. Though details of the modeling efforts are not

within the scope of this work, some comparisons are made between the modeled and

experimental results.
































Figure 1-1. Block diagram ofHPRTE flow path.









CHAPTER 2
LITERATURE REVIEW

History of Semi-Closed Cycle Concepts

Innovations in the Brayton cycle have continued to improve the performance of gas turbine

engines since their inception. One such improve, and perhaps the most pertinent to this work is

the semi-closure of the Brayton cycle. Although little development has been done recently, the

concept of semi-closure is nearly as old as gas turbines, as summarized by Gasparovic [1]. As

early as 1940s, attempts were made at developing semi-closed gas turbine engines. The initial

attempts at recirculating exhaust gases in a semi-closed cycle were plagued with corrosion

associated with the heavy fuel oils of the time. Subsequent efforts only recirculated air in a

secondary, "charging" circuit that added excessive complications and costs to the cycle.

In the late 1960s Gasparovic attempted to revive the previously abandoned idea of semi-

closer, as fuel technology had progressed. He also presented the idea of semi-closure

implemented with a low-pressure turbocharging system. In [1], he notes the potential of gas

turbine engines of a given power to be made smaller, estimating a pound-per-horsepower of

about 2/3 that of conventional engines. He also points out the advantage of part-load efficiency

achieved by varying the low-pressure charging of the cycle. Specifically, he cites the Sulzer

Brothers' design for a 20 MW plant operating at 32% efficiency at full power, and at 28%

efficiency at only half power. Again, this engine eventually failed due to excessive corrosion.

In recent years several research efforts at the University of Florida have once again

revitalized the semi-closed, turbocharged, gas turbine engine referred to as the HPRTE. These

studies have focused on many of the specific HPRTE attributes, as well as on various

applications thought to be well suited by this cycle. However, most of the more recent,

University of Florida studies were modeling analyses.










Analytical Studies

Nemec [2] conducted studies on the HPRTE with a Rankine bottoming cycle. In his work,

several bottoming cycle fluids were considered, as well as feedwater heating, superheating and

recuperation in relation to optimizing the overall thermal efficiency. One apparent challenge in

Nemec' s analysis was matching the bottoming-cycle performance with that of the topping cycle.

However, with the fluids and components considered, an optimal thermal efficiency of 54.5%

was reported. Nemec notes that higher efficiencies are possible with the inclusion of different

bottoming fluids and combined cycle components.

Landon [3] and Danias [4] performed several independent, but similar analytical studies on

the HPRTE. Landon's analysis evaluated the suitability of several HPRTE configurations for

marine applications and compared them with currently employed technology. Those results

indicate significantly higher part-load efficiencies, resulting in a 24% increase in range for a

given marine vessel. Danias' analysis also revealed similar qualities from two HPRTE cycles,

but for helicopter applications. His model predicted an increase in non-dimensional range of up

to 46% over a conventional helicopter engine resulting from the improved efficiency at off-

design power levels. Danias' simulations also predicted design-point efficiency increases of 30-

35% over the baseline engine, at the expense of a bearing a slightly larger and heavier

recuperated engine.

MacFarlane [5] investigated the implications of extracting water from the HPRTE

recirculation flow path. He also looked into any potential benefits of water re-inj section into the

cycle. His results show a maximum possible decrease of Specific Fuel Consumption (SFC) of

about 8%. Furthermore, MacFarlane found that water extraction could also increase the specific

power by about 4.5% over baseline. Re-inj section of water back into the cycle seemed to increase

SFC by about 2%, with little or no other impact on other performance parameters. However, it is









noted that the increase in SFC comes with the inj section of water in its liquid phase, and that this

increase would be less if the water were inj ected as steam.

Muley [6] carried out studies on the emissions-related effects of semi-closure. In his work,

Muley focused primarily on how the formation of thermal NOx was impacted by recirculating

exhaust gases. He discovered that an impressive reduction of thermal NOx could be realized by

semi-closure of a gas turbine power cycle. This effect grew more pronounced with further

dilution of the combustor inlet gases, yielding a decrease in thermal NOx of about eight to nine

orders of magnitude, depending also on the combustor inlet temperature.

Boza [7] was the first to perform a modeling analysis on the HPRTE power cycle

combined with a VARS in 2003. In his modeling efforts, Boza uncovered some striking

performance characteristics made possible by the combination of these cycles. He simulated a

large engine (40 MW) and a small engine (100kW), and evaluated their performance when

coupled to a Lithium-Bromide (Li-Br) VARS, considering both power and refrigeration as

beneficial outputs. Auxiliary refrigeration was presented as a percentage of nominal power. For

the large engine, Boza estimated a thermal efficiency of about 62% with 25% auxiliary

refrigeration at 850F ambient conditions. This efficiency dropped only two points for the same

operating point on a 1030F day. Boza points out that even higher efficiencies are calculated if

the auxiliary refrigeration is considered alongside of power in the efficiency definition. For the

small engine Boza showed a thermal efficiency of about 43%, but with an additional 50% of

auxiliary refrigeration, again on an 850F day. The thermal efficiency of this engine dropped

three points for the 1030F day. Boza also indicated that ambient conditions directly affected only

the VARS performance. This highlights the combined-cycle feature that refrigeration and









efficiency can be interchangeable, and that auxiliary refrigeration can be shunted back into the

HPRTE cycle to recover efficiency points if desired.

The most recent modeling work on the combined HPRTE/VARS cycle was completed by

Khan [8] in 2006. Khan's model was intended to predict and optimize the design-point

performance of a medium-sized engine with conservative, but modern efficiencies and

temperature limitations. Khan's models added to Boza's work in that he accounted for the water

condensation associated with the gas-path cooling, and explored the implications of using

different refrigerants. The figure of merit used was a linear combination of efficiency, auxiliary

refrigeration, and water extraction. His model predicted an attainable thermal efficiency of

40.5% at an optimal low-pressure-spool pressure ratio of two, and a turbine inlet temperature of

around 25500F. Khan also showed that 1.5 pounds of water could be extracted from the cycle

for every pound of propane fuel consumed. Khan' s results are similar to Boza' s small-engine

results with the exception of his predicted external refrigeration--38% of nominal power for a

Li-Br cycle. This difference is probably related to Khan's accounting for water condensation.

Experimental Studies

Several experimental research initiatives have also been carried out at the University of

Florida to further explore characteristics of HPRTE performance and operation. One program in

particular was funded by NASA in the late 1990s to evaluate a turbocharged Titan T62T32A gas

turbine engine operating in a semi-closed cycle configuration. This program [9] proved many of

the concepts upon which the HPRTE is founded. Though plagued by hardware problems

unrelated to the focus of the research, HPRTE operations yielded power increases of up to 70%

by turbocharging. Shortcomings in the fuel and control systems limited this test program from

fully realizing the constant efficiency and SFC potentials of the HPRTE. However, some simple

scaling arguments provided within [9] show that with more accommodating controls and fuel










components, the efficiency and SFC would remain much more constant for various operating

points. The anticipated, positive impact of semi-closure on emissions also was validated within

this test program. The HPRTE CO emissions were observed to be a factor of 25 less than the

baseline values. This was thought to result from larger proportions of water in the working fluid

due to recirculation. Spray coolers were also employed on the apparatus to aid in recirculation-

gas cooling, and also contributed to a higher water concentration in the working fluid. The

reduction of NOx emissions was also observed, but to a lesser extent than predicted by Muley in

[6]. The potential for reduced NOx emissions was hindered by lower-than-expected recirculation

flow rates and non-uniformities in the combustor temperature field. Other experimental efforts,

detailed in [10] and also funded by NASA, focused on the development of more suitable

combustors for stable, semi-closed engine operations.

Previous experimental efforts have, by necessity, incorporated some means of cooling to

the recirculation gas path, as this is essential for reliable and efficient engine operation. Most

frequently, one or more air-to-water intercoolers were employed to meet this need. To date,

based on the apparent absence of any relevant literature, it seems as if no experiments have been

performed on a semi-closed cycle gas turbine combined with a VARS, so the present study

represents the first of its kind.









CHAPTER 3
TESTING HARDWARE

An experimental apparatus was conceived for demonstrating the capabilities of the

HPRTE/VARS combined cycle. This apparatus was centered around a Rover 1S-60 gas turbine

engine and a custom-built VARS. In order to operate the combined cycle and successfully

acquire useful data, a host of other components were necessary, such as facility cooling resources

and instrumentation. This chapter describes each of the elements in the HPRTE/VARS

combined cycle, their roles, and their specifications, where available.

Combined Cycle Components

This section lists the various components relating to the combined cycle, along with a

detailed description for each.

Engine Characteristics and Specifications

The Rover 1S-60 is a single spool turboshaft engine that was designed and manufactured

in the late 1950s and early 1960s. It was primarily used for Auxiliary Power Units (APU) and

fire-fighting water pumps before being used as an educational tool. The engine remained in

production for about 20 years during which its safety, reliability, and longevity were well proven

[1l].

The engine utilizes a 19 blade, radial compressor and radial diffuser vanes with a design

flow rate of about 1.33 lbs/s (0.603 kg/s), and pressure ratio of 2.8: 1. To accommodate the dilute

oxygen concentrations in the semi-closed cycle, the reverse-flow combustor was modified by

welding stainless steel straps over some of the dilution holes, forcing more gases--and more

oxygen--into the primary zone. Modiaications were also made to intercept the flow from the

compressor, redirecting it through the recuperator before returning it to the combustor. The

working fluid is then expanded across a 30-blade axial turbine. The nominal rated output of this










engine is 60 hp (45 kW) at a mechanically governed shaft speed of 46,000 rpm, with a thermal

efficiency around 14%. The rotor speed is reduced in a gearbox on the front end of the engine to

3000 rpm, where a shaft transmits power to a water-brake dynamometer.

Further modifications were also made to the engine casing, replacing it with thicker steel.

This was necessary before pressurized operations could be safely attempted since the pressure

within the casing was expected to increase by as much as a factor of two.

Recuperator

The recuperator is a custom unit designed and fabricated by Elanco. It is a single-pass

tube-and-shell heat exchanger with 672, 0.375-inch (9.53 mm) stainless steel tubes extending 27

inches (69 cm) through the shell-side. The tubes are enclosed with 14-inch (36 cm) diameter, 40

gauge stainless steel [12]. The hot-side inlet is attached to the turbine exhaust ducting with a V-

band clamp, while the hot-side uses an 8-bolt flange to connect to downstream ducting. The

cold-side ducting attaches with 12-bolt flanges. The recuperator design effectiveness is about

0.51 [12].

Turbocharger

The turbocharger used was a Garret GT 4294-731376-1, with a turbine scroll Area to

Radius (AR) ratio of 1.44. This turbocharger was chosen primarily for its superb compressor

efficiency, which would delivery higher pressure ratios with relatively cooler turbine-side inlet

temperatures. Additionally, the Garret 4294 was shown by an off-design model to match well

with desired HPRTE operating points [13]. The turbocharger relies on an independent oil pump,

dry sump, and oil cooler to provide its lubrication requirements.

Heat Exchangers

Three heat exchangers were made necessary by the recirculation of hot exhaust gases to the

front end of the engine. All three are arranged in series in the recirculation line, and each is









named according to the qualitative state of the recirculation gases entering it: the Hot Gas

Cooler (HGC), Warm Gas Cooler (WGC), and Cold Gas Cooler (CGC).

Hot gas cooler (HGC)

The HGC is a custom heat exchanger designed and fabricated by Energy Concepts in

Annapolis Maryland for this application. End-to-end, it is 36 inches (91 cm) long, and

transitions from six inch Outside Diameter (OD) ducting, to a 12 inch (30 cm) OD shell, and

back to the six inch ducting. No baffles are employed within the heat exchanger, and all

stainless steel construction is used. The tube-side heat transfer medium is a strong solution of

ammonium and water. There is intended to be phase change within the HGC tube-side, since it

serves as the vapor generator for the VARS.

Warm gas cooler (WGC)

The WGC was necessary if sub-ambient inlet temperatures were to be possible. It serves

the purpose of rej ecting heat to ambient (though it uses a cooler-than-ambient medium) before

recirculating gases are last cooled by the CGC. The WGC is an air-to-water heat exchanger that

makes use of the facilities chilled water circuit. Chilled water makes two passes in the tube-side,

while the recirculation gases turn through several baffles on one shell pass. Again, Elanco

designed and manufactured the WGC using all stainless steel construction. The design

effectiveness of this heat exchanger is 0.85 [12].

Cold gas cooler (CGC)

The CGC was also constructed of stainless steel by Energy Concepts, and functioned as the

evaporator for the VARS. This heat exchanger is, in principle, capable of chilling the High

Pressure Compressor (HPC) inlet gases to sub-ambient temperatures. It is longer than the

HGC-about 48 inches (1.2 m)--with the same types of transitions on each end, and also uses

no baffles within. Ideally, the heat transfer medium is pure, liquid ammonia undergoing phase









change with a few degrees of superheat upon exiting the CGC. The gases on the shell side

consist of recirculation gases mixed with fresh air from the Low Pressure Compressor (LPC).

Dynamometer

The dynamometer is contemporary with the Rover 1S-60 and was sold as an integral

package with the engine. It is a water-brake manufactured by Heenan & Froude (now Froude

Hofmann) and is equipped with a mechanical tachometer. The spring balance with which the

dynamometer was originally equipped was replaced with a load cell and reader. The original,

mechanical tachometer is also supplemented with an optical tachometer directed at reflective

tape on the Power Take-Off (PTO).

Ducting

The ducting used throughout the HPRTE rig was six-inch (15 cm) OD aluminized steel

with a thickness of 0.070 inches (1.78 mm). It was manufactured for use as exhaust pipe for

large trucks, and was chosen for its availability, low cost, and light weight. Pipe-to-pipe

connections were made using 0.375 inch (9.53 cm) thick, 12-bolt custom flanges which were

welded to each pipe-end.

Facility Resources

There were several of the building assets that were both useful and essential for conducting

engine runs. These dealt primarily with the issue of removing heat from the engine room. The

building chilled water circuit served to remove heat from the power and refrigeration cycles,

while the ventilation systems available helped to maintain uniform temperatures within the

EGDSL during engine runs.

Chilled Water

The building in which the EGDSL is situated utilizes a closed-circuit Process Chilled

Water (PCW) system to function as a cooling medium for several labs in the building. This










system was tapped into and can be accessed from one of two headers in the EGDSL: one on the

north wall, and one on the south wall. Each header is capable of delivering about 35 gpm (2.2

kg/s), or one header can deliver about 64 gpm (4.0 kg/s). All of the heat rejected from the

recirculation gas path ultimately finds its way into this cooling circuit. The VARS is cooled by

the south header, and the WGC by the north header. The PCW circuit in turn rej ects its heat to a

stacked plate heat exchanger elsewhere in the building, which has a maximum heat transfer rate

of almost 200 TR (703 kW). PCW temperatures are typically controlled to be around 55 OF (286

K), but by altering the PCW control scheme can be made as cool as 42 OF (279 K).

Fans and HVAC

To impede the accumulation of harmful exhaust fumes within the testing facility, two fans

have been employed to move fresh air through the EGDSL. An existing air conditioning unit

also exists in the EGDSL, the service of which is shared with a neighboring room. While this

unit does little to cool the room during engine operations, it does help to move more fresh air

from elsewhere into the EGDSL.

Vapor Absorption Refrigeration System (VARS)

The VARS is a custom refrigeration unit designed and built by Energy Concepts. The

VARS is a unique, single-effect, ammonia-water absorption refrigeration system with a

maximum evaporator heat load of 19 TR (67 kW), at a COP of 0.85. The VARS interacts with

the HPRTE via the HGC and CGC described above, and rej ects heat from the absorber and

condenser through the south PCW header.

Overview of Build Configurations

To help ensure that all of the hypothesized attributes of this combined cycle could be

demonstrated and explored, several small, progressive steps were taken with the experimental

apparatus. With some changes in the apparatus different components were chosen and










implemented, and with other changes the same components were used, but in a different

configuration. The configurations generally progressed such that different capabilities were

accentuated with different configurations, ultimately arriving at an apparatus capable of

exploring all of the HPRTE/VARS combined cycle attributes. There were four different build

configurations: Build 1, Build 2A, Build 3, and Build 4.

Build 1

Build 1 was one of the first HPRTE testing platforms utilized in this series of tests. It was

the configuration used for some of the previous research work discussed in Chapter 2. A block

diagram of Build 1 is shown in Figure 3-1.

The Build 1 configuration initially relied on a single, air-to-water heat exchanger to reject

all of the heat from the recirculated gas line. This is the HGC shown in Figure 3-1.

Subsequently, a surrogate VARS evaporator was added to the recirculation flow path to

investigate the effects of additional, low-temperature recirculation cooling. Also, new water

extraction hardware was designed and built to accommodate the newly addition condensation

location. The evaporator was part of a small, vapor compression refrigeration system capable of

yielding about one ton (3.5 kW) of cooling. It was installed within a 0.250-inch (6.35 mm) thick

steel box that previously served to house a filter. A photograph of this installation is shown in

Figure 3-2.

The improvised evaporator yielded limited success in furthering the water extraction for

Build 1. However, an important point was made clearer. Significant water extraction could still

be attained without cooling all of the recirculated gas mixture down to its saturation temperature.

The cold tubes accomplished the local cooling of gases near to the tube surfaces, and

successfully condensed appreciable quantities of water.









Overall, Build 1 tests were successful in proving the concept of water extraction from the

HPRTE. However, no efforts were made to pressurize this engine configuration.

Build 2A

It was believed that the second phase of this research would permit pressurized engine

operations, and greater gas-path cooling. Before moving prematurely to a ducting configuration

incorporating a VARS, more simulated VARS cooling was used instead. This decision was

influenced by the availability of two air-to-water heat exchangers in the EGDSL. These could be

implemented practically, allowing more to be learned about recirculation cooling in a reasonable

amount of time. This successive ducting configuration was referred to as Build 2A. A block

diagram of the Build 2A configuration is provided in Figure 3-3.

Build 2A utilized two, tube-and-shell heat exchangers to reject gas-path heat to the

building PCW circuit. A summative 40-50 TR (140-175 kW) was consistently removed from the

recirculation line with these two coolers. This helped to demonstrate some of the concepts

related to VARS integration since the means by which the recirculation gases are cooled are

arbitrary. These two heat exchangers simulated the VARS vapor generator and evaporator to

some extent, though the cooling medium (PCW) was different.

In addition to yielding exceptional water extraction results, the Build 2A configuration

provided some valuable information about pressurized operations. It was observed that some

means of throttling is necessary between the LPT inlet and LPC exit in order to pressurize.

Otherwise, the exhaust gases may flow the wrong way through the LPC, out to ambient, since,

depending on the ducting configuration, this can be the less resistive path. Having this fact

reinforced experimentally was found useful when considering the next build configuration.









Build 3

The next build configuration was called Build 3. In this configuration, one of the air-to-

water heat exchangers was replaced with the vapor generator of the VARS. Also, one more heat

exchanger, the CGC, was added to the recirculation line. The CGC is the evaporator in the

VARS and provides the opportunity to condense more water than the previous build with a lower

temperature heat transfer medium. Furthermore, Build 3 was the first configuration to utilize the

new, modified Rover 1S-60 engine. The new engine was equipped with a much thicker casing

made necessary by higher pressure operations. It was feared that the Rover 1S-60, as originally

manufactured, would have ruptured with the greater HPC exit pressures within. Lastly, the Build

3 configuration benefited from a new Garret 4294 turbocharger. This turbocharger was chosen

because it was more capable of higher pressure ratios, given the same, relatively cool recuperator

exit temperature. With the previous turbocharger, models showed that pressure ratios of 2:1

were unattainable with the relatively low LPT inlet temperature. Plans were originally in place

to duct some of the hot, High Pressure Turbine (HPT) exit gases around the recuperator, but the

replacement of the old turbocharger with newer technology was more practical. A block

diagram of the Build 3 configuration is shown in Figure 3-4.

One characteristic of the Build 3 configuration was that the LPC exit gases mixed with the

recirculation gases immediately downstream of the CGC. This was beneficial in that the

pressure drop between the LPC exit and HPC inlet was minimal. However, the drawback to this

mixing location was that the LPC exit flow was cooled only by mixing with the cooler

recirculation gases. At higher LPC pressure ratios, and LPC exit temperatures, the mixed

temperature of the HPC inlet gases would be approaching the limit imposed by the manufacturer.

In consequence, a move to the final configuration was made.









Build 4

The final configuration, Build 4, involved only a small change in the gas path. The

abovementioned LPC exit mixing location was moved to an elbow directly upstream of the

CGC. By so doing, the mixed recirculated exhaust gases and fresh LPC air would all pass

through the CGC together to be cooled. Initially, this may seem to be no different, in a

thermodynamic sense, as the same quantity of heat is removed from the gas path whether mixing

is done before or after the CGC. However, by mixing before the CGC, the flow rate, and

Reynolds number passing through it increased from 60% to 100%, depending on the

recirculation ratio and LPC flow rate. Moreover, the CGC inlet temperature will generally be

increased by the mixing, since the LPC exit temperatures were usually higher than the WGC exit

temperatures. These two effects, higher Reynolds number and higher inlet temperature, were

expected to increase the effectiveness of the CGC. Thus, for the same conditions, the HPC inlet

temperatures were expected to be lower for Build 4 than Build 3. A block diagram of Build 4 is

shown in Figure 3-5.

Data Measurement

Testing of the HPRTE/VARS combined cycle necessitated the use of various types of

instrumentation. This section describes in detail all of the hardware and software used to

measure and acquire experimental measurements.

Data Acquisition Hardware

The Data AcQuisition (DAQ) hardware is here defined as the equipment between all

transducer wire terminations and the data output. Hence, there are two components that provide

this throughput of information: a National Instruments chassis, and a personal computer. The

purpose of the DAQ hardware is to receive the raw, analog signals from various transducers and,

after filtering and amplification, convert each of them into a digital signal. From this point a









1200 MHz Personal Computer (PC) operating with 256 Mb of RAM received the digital stream

and readied it for the DAQ software.

The DAQ hardware used for these experiments consisted of one chassis, four modules and

four terminal blocks. The chassis is a model SCXI 1000, and provided a low-noise environment

in which signal conditioning can be performed by the modules. A 50-pin cable conveyed the

data to a PCMCIA card installed within the PC.

Two SCXI 1 102 modules were employed for the acquisition of thermocouple signals--one

for the HPRTE temperatures, and one for the VARS temperatures. The SCXI 1102 modules

were equipped with 32 differential, analog channel inputs, not including one Cold-Junction-

Compensation (CJC) channel. Additionally, each channel was amplified with a gain of either

one or 100 at the discretion of the user. Each channel in the 1 102 modules was also equipped

with a three-pole lowpass filter rejecting 60Hz noise. SCXI 1300 terminal blocks were used with

each of the SCXI 1 102 modules for the thermocouple wire terminations.

One SCXI 1 100 module was utilized to condition all of the pressure transducer, load cell,

and tachometer reader signals. The SCXI 1100 also had 32 differential channels and an onboard

programmable gain instrument amplifier. All channels were also equipped with a jumper-

selectable (four Hz or ten k
in a SCXI 1303 terminal block.

One SCXI 1 126 module was necessary for the acquisition of all instrumentation having a

frequency output. It was equipped with eight isolated channels with filtering, and a software

programmable frequency-to-voltage conversion circuit. This module used a SCXI 1327 terminal

block for its wire terminations. Within this terminal block was a selectable 1:1 or 100: 1 voltage

attenuation switch for each channel.










Data Acquisition Software

The software used for processing the data signals was LabVIEW 7.1. It is also a National

Instruments product, and therefore interfaces very well with the DAQ hardware. LabVIEW is a

graphical programming environment that allows instrumentation signals to be monitored,

manipulated or scaled, and appended to a data file, all in real-time. This type of interface was

essential to monitoring key parameters on both the power and refrigeration sides of the combined

cycle.

Instrumentation

In general, five types of measurements were made in testing the HPRTE/VARS:

temperature, pressure, flow rate, rotational speed, and load. Thermocouples were exclusively

used for temperature measurements. For pressure measurements, diaphragm-type pressure

transducers were implemented to provide the DAQ with an analog signal. In addition to these,

redundant manometers and mechanical pressure gauges were used for some of the more vital

pressure measurements. Where measured directly, liquid flow rates were acquired by either

turbine or paddle-wheel, self-powered flowmeters. The only rotational speed measurement,

dynamometer speed, was made with an optical tachometer, and load measurements were

performed with load cells.

Thermocouples

Three different types of thermocouples were used for temperature measurements: J-type,

K-type, and T-type. In general, the K-type, J-type, and T-type thermocouples measured hot gas,

warm gas, and cool liquid temperatures, respectively. Exceptions to this are on the VARS,

where T-type thermocouples were used for all temperature measurements. Most of the

thermocouples were 0.250 inches (6.35 mm) in diameter, and extended to approximately the

centerline of the flow path.









Pressure transducers

The pressure transducers used were all Omega PX138 series sensors. These transducers

employed an eight VDC power source with a one VDC to six VDC output, and were equipped

with temperature compensation up to 122 OF (323 K). There were four types of differential

PX138 sensors used, classified by pressure range: 0-100 psi (689 kPa), 0-30 psi (207 kPa), 0-5

psi (34 kPa), and 0-1 psi (7 kPa). All of these sensors were installed near the ceiling of the

EGDSL, where shielded cables then carried their signals to the SCXI 1303 terminal block.

Turbine and paddle-wheel flowmeters

The fuel flow rate measurements were made using a Hoffer Flow Controls, low-flow liquid

turbine flow meter, model number IVFl1/2X70B. It was situated immediately downstream of the

fuel tank and generated a frequency output in proportion to the volumetric flow rate. For a

constant viscosity, this flowmeter establishes a linear response, which is transmitted directly into

the SCXI 1327 terminal block. After correcting for an initial offset, accuracies within around

1% are typical from this instrument.

Two paddle-wheel flow meters were used to measure the volumetric flow rate of PCW-

one at each header in the EGDSL. These units were manufactured by Omega Engineering,

model number FP-5300. These instruments came with CPVC pipe fittings that would ensure the

proper number of pipe diameters were in place, both upstream and down stream of the

flowmeters. These flowmeters also generated a linear frequency output over volumetric flow

rate.

Optical tachometer

Dynamometer speed measurements were accomplished with a Monarch ROS-5W optical

sensor and ACT-3 reader. The sensor was placed about 0.5 inches (1.3 cm) from the PTO, the

circumference of which was covered with electrical tape. A single strip of reflective tape was










placed over the electrical tape, triggering the optical sensor once per dynamometer revolution.

The reader displayed the dynamometer speed in the control room, but also provided a 0-5 VDC

output, from which the DAQ hardware received its speed signal.

Load cells

Two load cells were used during engine tests. The more important of the two was used to

counter the dynamometer torque. This instrument was necessary for measuring the engine load,

and in turn the net power output. The output from this load cell was interpreted and displayed by

an Omega Engineering DP-145 multi-purpose reader. Unfortunately, the specifications for this

particular load cell could not be located.

The second load cell was employed to measure the quantity of water being extracted from

the HPRTE in real-time. A reservoir was suspended from this instrument, which was

cantilevered from a small support structure. Several hoses conveyed the condensate from within

the ducting into the reservoir over which they were suspended, also from the support structure.

Thus, the real-time weight of the extracted water could be straightforwardly differentiated to

obtain a flow rate. This particular load cell was an Omega Engineering platform load cell with a

72 lb (33 kg)capacity.

Gas analysis

A Cosa 1600 IR analyzer was used to evaluate the composition of exhaust gas samples

taken from the exhaust before leaving the EGDSL. It is a portable, five-gas exhaust gas

analyzer. It quantifies carbon monoxide (CO), carbon dioxide (CO2), and unburned hydrocarbon

(UBHC) concentrations using non-dispersive infrared technology. Diatomic oxygen (Oz) and

nitric oxide (NO) concentrations can also be measured using electrochemical sensors.










Mechanical instruments and analog readers

The HPRTE/VARS experiments did not rely exclusively on the DAQ system for all

values, but also utilized redundant sets of manually recorded data. This practice was adopted

primarily to insure against the complete loss of data in the event of digital data corruption.

Another benefit from practicing manual data recording stems from the fact that much of these

data, particularly the pressures and differential pressures, were measured with different

mechanical instruments. In these cases, the analog and digital sets of values could be compared

to mutually ensure the reliability of both. However, this wasn't the case for the temperature

measurements since the signal from each thermocouple bifurcated to an analog reader and the

DAQ. Consequently, both values were always the same.

Each type of thermocouple had its own multiplexer and Omega DP460 reader. Thus, three

multiplexer/reader pairs were implemented to display all of the HPRTE and VARS temperatures.

Specifics on each of the temperature measurement locations and their respective thermocouples

are provided in Table 3-1. The state points correspond with those shown in Figure 1-1.

All of the HPRTE pressure taps branched off to a redundant gauge of some sort. The more

important of these pressures, deemed vital to engine health during experiments, were routed to

their own individual gauge. This improved the visibility of important pressures. Other, less

critical pressure hoses were each connected to one of two headers through its own valve. Each

header was connected to a different Bourdon tube pressure gauge. The pressure panel operator

could choose to view a specific pressure by simply opening the valve for that hose. After

viewing and recording a pressure, the panel operator would then close the valve, and vent the

header. The header would then be ready to receive a new pressure. Details on each of the state

point gauges are shown in Table 3-2.









All of the differential pressures were treated similarly. Each differential pressure line

going to a transducer also teed off to some sort of manometer. Manometers of various sizes and

shapes were used, employing sundry fluids. Table 3-2 also presents the specifications for each

of the manometers used, alongside the specific differential pressures for which they were used.

Lastly, the fuel flow rate utilized an additional rotameter, inline with the turbine flow

meter. This device was always viewed helped validate the weak, self-powered signal of the

turbine flowmeter. The rotameter outputs a mass flow rate, and is calibrated for a specific

gravity of 0.835.

Table 3-1. Temperature locations and instrumentation.
IExpected Range
State Point Location Thermocouple Type Instrument Range Vle


-40-13800F
(251-1022 K)
-40-13800F
(251-1022 K)
-40-13800F
(251-1022 K)
-40-13800F
(251-1022 K)
-40-13800F
(251-1022 K)
-40-13800F
(251-1022 K)
-40-13800F
(251-1022 K)
-40-13800F
(251-1022 K)
-320-22000F
(77-1478 K)
-320-22000F
(77-1478 K)
-320-22000F
(77-1478 K)
-320-22000F
(77-1478 K)
-40-13800F
(251-1022 K)


40-1000F
(278-311 ED
40-1100F
(278-3161{)
100-2500F
(311-394 K)
40-1200F
(278-322 K)
40-1200F
(278-322 K)
300-3800F
(422-4661{)
350-4200F
(450-4891{)
580-6200F
(578-605 ED
1300-16000F
(978-1144 K)
900-12000F
(755-922 K)
780-11000F
(689-866 K)
780-11000F
(689-866 K)
500-6000F
(533-589 K)


Ambient

LPC inlet

LPC exit

CGC exit

HPC inlet

HPC exit

HPR inlet

HPR exit

CO1VB exit

HPT exit

LPR exit

LPT inlet

WGC inlet










Expected Range
of Values
110-1850F
(3 16-3 58 K)
800-9000F
(700-755 K)
42-750F
(279-297 K)
42-850F
(279-303 K)
42-750F
(279-297 K)
42-850F
(279-303 K)


State Point


Location


Thermocouple Type


Instrument Range
-40-13800F
(251-1022 K)
-320-22000F
(77-1478 K)
-320-6600F
(77-622 K)
-320-6600F
(77-622 K)
-320-6600F
(77-622 K)
-320-6600F
(77-622 K)


WGC exit J

LPT exit K


PCW supply

PCW return

PCW supply

PCW return


Table 3-2. Pressure locations and instrumentation.


Expected Range
of Values

736-788 mm-Hg

0-2 in-H20
(0-4 mm-Hg)
0-7 in-H20
(0-13 mm-Hg)
0-15 psig
(0-103 kPa)
0-48 in-H20
(0-90 mm-Hg)
-0.5-15 psig
(-3.4-103 kPa)
0-65 psig
(0-448 kPa)
0-65 psig
(0-448 kPa)
0-50 in-oil
(0-164 mm-Hg)
0-60 psig
(0-414 kPa)
0-20 psig
(0-138 kPa)
0-20 psig
(0-138 kPa)
0-20 psig
(0-138 kPa)


State Point


Location

Ambient


Type of Instrument

Internet Resource

Angle Manometer
SG 0.827
Angle Manometer
SG 1.91

Bourdon Tube Gauge
Manometer
SG 1.0


Instrument Range


N/A


0-2 in-H20
(0-4 mm-Hg)
0-7 in-H20
(0-13 mm-Hg)
0-60 psig
(0-414 kPa)
0-60 in-H20
(0-112 mm-Hg)
0-60 psig
(0-414 kPa)
0-100 psig
(0-689 kPa)
0-200 psig
(0-1379 kPa)
0-60 in-oil
(0-197 mm-Hg)
0-200 psig
(0-1379 kPa)
0-60 psig
(0-414 kPa)
0-200 psig
(0-1379 kPa)
0-200 psig
(0-1379 kPa)


LPC inlet dP

MAI dP

LPC exit

CGC dP


0-2B


2.8-2.9


HPC inlet Bourdon Tube Gauge

HPC exit Bourdon Tube Gauge

HPR inlet Bourdon Tube Gauge


Manometer
SG 1.75


COMB dP


HPT inlet Bourdon Tube Gauge


HPT exit

LPR exit

WGC inlet


Bourdon Tube Gauge

Bourdon Tube Gauge

Bourdon Tube Gauge










Expected Range
Instrument Range
of Values
0-70 in-oil 0-65 in-oil
(0-109 mm-Hg) (0-101 mm-Hg)
0-200 psig 0-20psig
(0-1379 kPa) (0-138 kPa)


State Point


Location

RCV dP

LPT exit


Type of Instrument
U-Tube Manometer
SG 0.827

Bourdon Tube Gauge


Figure 3-1. Block diagram of Build 1 configuration.
































Figure 3-2. Photograph of Build 1 improvised evaporator installation.


Figure 3-3. Block diagram of Build 2A configuration.































Figure 3-4. Block diagram of Build 3 configuration.


Figure 3-5. Block diagram of Build 4 configuration.









CHAPTER 4
TEST PROCEDURES

Test Plan

Before conducting an engine test, a test plan needed to be prepared so that time and fuel

were not wasted making decisions during the run. Fist, the test plan identified the operating

points to be evaluated. Then a plan was conceived to achieve each operating point since

transitions into the pressurized regime were only successful after taking the proper steps in the

correct order. Such a plan usually consisted of the gradual closing of the boost control valve

(VEXH in Figure 1-1), while carefully maintaining a reasonable recirculation ratio and fuel-to-air

ratio using the recirculation valve (VREC in Figure 1-1).

Contingency plans were also needed in case a particular operating point could not be

reached. It was often the case, for example, that a desired pressure ratio could not be reached

using the test plan. Instead of shutting down the engine and allowing the day's efforts to go to

waste, it was deemed prudent to have a secondary obj ective for the same engine run. An

example of some operating procedures is provided in Appendix A.

Engine/Component Preparation

Prior to starting the HPRTE and VARS, it was essential that several, pre-planned steps be

taken to ensure personnel safety, engine longevity, and data integrity. A thorough checklist was

created for each engine run, and rigorously reviewed before that run. Quantities such as oil

level, battery charge, PCW flow rate, etc. were confirmed as acceptable. Following these

preparations, a regular startup procedure was employed in an attempt to further remove any

question regarding the consistency and integrity of all engine runs. An example of these setup

procedures is provided in Appendix B.










Instrument Preparation

Additional actions were taken before engine runs specifically to prepare the

instrumentation, DAQ hardware, and DAQ software. Some checks were made before every

engine run, while others were less frequent. However, all were deemed necessary for acquiring

consistent, accurate, and controlled data from the HPRTE/VARS combined cycle.

Transducer Offsets

Since the DAQ hardware was equipped with CJC capabilities, there was little maintenance

necessary for the thermocouple measurements. However, infrequent and random tests were

employed to make certain that thermocouples were maintaining their accuracy and precision.

Thermocouples were immersed in a small pool of boiling water away from the heated surface to

ascertain whether or not the proper boiling temperature was being output. Thermocouples rarely

deviated more than 1 OF (0.6 K) or so from the expected temperature. If one particular

thermocouple was somewhat off at this temperature, then an offset was added to that signal

within the DAQ software. Such a simple assessment may be flawed in that it is only performed

at one temperature in the thermocouple range. It was assumed that the thermocouple linearity

was retained, since calibrations at higher temperatures were not practical.

The pressure transducers also required that measures be taken to ensure consistent outputs.

Periodically, the transducer calibrations were checked to verify that at least their pressure-over-

voltage slope had remained the same. Any offsets in their outputs were addressed at the

conclusion of engine runs, but before analysis of the data.

The linear calibration line frequently shifted up or down by a small amount. To correct for

this several minutes of lead-in data were recorded before each engine startup. During the lead-in

interval, all of the pressure transducers should have indicated a gauge pressure of zero. The lead-

in data was averaged, and then subtracted from all subsequent data points for that engine run.









The intended result was that each pressure measured as zero psig (0 kPa) before startup and

following shutdown.

The pressure transducers were powered with a variable output DC power supply. All

transducer calibrations were performed with an excitation voltage of 8.11 VDC. Any drifts in

this output were compensated for within the DAQ software. Before each raw voltage output was

translated to a pressure, the signal was scaled according to the real-time excitation voltage (also

read into the DAQ) relative to the calibration voltage. This effectively compensated and drift in

the excitation voltage, since the transducer output voltages scale one-to-one with the excitation

voltage.

Ambient Conditions

The performance of any gas turbine engine depends greatly on the thermodynamic

conditions of its environment. As such, the ambient temperature, pressure and humidity were

measured or calculated. The ambient temperature was obtained from one of two J-type

thermocouples. The first of these was mounted in front of the Main Air Inlet (MAI) valve (V1\fAI

in Figure 1-1). However, for most of the Build 4 engine runs this valve was closed and none of

the HPRTE working fluid moved over that thermocouple. Thus, the second of these

thermocouples, mounted in front of the LPC inlet bellmouth, was most often used to represent

the ambient temperature. The two thermocouples usually measured different temperatures

because of their locations. This resulted from the way in which air was drawn through the

EGDSL. Fresh air from outdoors moved over the MAI thermocouple first, and then flowed over

and around some of the hot ducting, heat exchangers, and engine surfaces before reaching the

LPC inlet. Accordingly, the LPC inlet flow was a few degrees warmer than the MAI flow.

No barometers were used to directly measure the atmospheric pressure in the EGDSL.

This value was obtained from the University of Florida Department of Physics Weather Station,









which publishes this data online. These values are updated every five minutes, and were

recorded at the beginning and end of an engine run. The Weather Station website also stores 24

hour trends in case knowledge of barometric pressure variations during the run is desired.

The wet and dry bulb temperatures were measured directly within the EGDSL using a

psychrometer. This measurement was repeated at the beginning of each trial during an engine

run. At the conclusion of each run these temperatures were used with a psychrometric chart to

obtain the humidity ratio and relative humidity for each trial.









CHAPTER 5
DATA REDUCTION SCHEME

During the four months of engine testing, millions of data were acquired, along with the

challenge of analyzing them. For the period of each engine run, there were 63 values being

measured and recorded, once every 1.5 to two seconds. Each of these 63 pressures, temperatures

and flow rates needed to averaged, corrected where necessary, and then processed to arrive at

performance parameters such as power, efficiency, heat load, etc. Thus, an algorithm was

developed within the Matlab environment to carry out all of the statistical operations, and the

reduction and organization of data. The methods employed to reduce the HPRTE data are

detailed below. However, those measurements obtained from the VARS instruments were used

exclusively for VARS operations, and were too few to suffice for a performance analysis.

Consequently, only the power cycle data were analyzed.

As with any engineering analyses, some assumptions and approximations were necessary

in order to have a complete set of equations. Since many of the desired performance parameters

cannot be measured directly, each thermodynamic process and substance needed to be classified

so that it would receive the proper treatment. Additionally, the geometric, fluid, and heat

transfer properties of various components were often too complicated for the application of some

analyses, further compelling the use of good engineering assumptions.

Gas Properties

Given the primary role played by the working fluid in a power cycle, it is necessary to

describe how these fluids are analytically treated. The HPRTE presents a unique flow path

wherein many various thermodynamic processes take place: mixing of gases, power transfer

to/from gases, heat transfer to/from gases, in addition to the condensation and subsequent









extraction of water. This section will describe the assumptions, approximations, and equations

specific to the HPRTE gases.

Enthalpy and Specific Heat

The working fluid within the power cycle was approximated as a thermally perfect gas,

composed of N2, Oz, CO2, water vapor (H20), and CO, and as behaving in accordance with the

perfect gas law shown in Equation 5.1.

P =pRwxT (5.1)

Here the mass-specific gas constant, RhIX, is computed with Equations 5.2, using knowledge (or

assumed knowledge) of the chemical composition at a giving state point:


Rwx (5.2a)


M. = [yllM,
(5.2b)

where yi represents the mole fraction of the ith species being considered, Mi represents the

molecular weight of the i.th species, and Mulx represents the molecular weight of the mixture.

The variable, R is the universal gas constant.

For none of the calculations was the working fluid approximated as calorically perfect.

Instead, a third-degree polynomial was used to obtain the constant-pressure specific heat (c,) of a

given gas at a given temperature. The system used for obtaining c, for one of the pure gases

comprising the mixture is shown in Equations 5.3 [14].


P=1+ar+br2 +C3 (5.3a)



r = 1 (5.3b)









In Equations 5.3, the coefficients, a, b, and c, are particular to the pure gas being considered, and

cpo is the constant pressure specific heat of the pure gas at the reference temperature (To) of

300K. In Equation 5.3b, T is the temperature of the pure gas, which is assumed to be the same as

the mixture temperature. Following the computation of each constituent c,, that of the gas

mixture could be calculated using Equation 5.4.


cP,MIX CY CP,lM, (5.4)


The above describes only how, when necessary, the c, was computed for a gas mixture at a

single, given temperature. However, the more pertinent and more frequently occurring

computations were the specific enthalpy changes (Ah) within the various components. The Ah

calculations were carried out directly by combining Equations 5.3 and integrating definitely over

the measured temperature difference, per Equation 5.5.

ro"'T 2 bTo T cTo T
Ah = c,dT =c T + a T + 1 +1(55
J po 2To 3 T 4 T


Here again, the Ah on the left-hand side of Equations 5.5 is for a single species within the gas

mixture, and needs to be computed for each of the five constituents. Finally, the Ah of the gas

mixture could be calculated using Equation 5.6.


Ah C = f yh, AhzM (5.6)


It was often necessary to make use of an average cP arCOss certain of the thermodynamic

processes. For example, some of the turbomachinery with large temperature differences needed

an average specific heat ratio to use in efficiency calculations. For cases in which a temperature

difference was measured, and Ah calculated, Equations 5.7 were used to obtain an average

constant pressure specific heat (cP,AVG), and specific heat ratio (YAVG).










Ah
cP,.4m (5.7a)


Turn (5.7b)

cP,ATU

Gas Composition

Specialized instrumentation was employed to measure on the composition of gases

throughout the HPRTE flow path. It was impractical, however, to directly measure gas samples

from more than one location, because of the time needed to acquire just one sample.

Consequently, assumptions were made allowing the gas composition to be followed through

various, HPRTE processes effecting changes in the gas mole fractions. One substance of chief

importance in this study was the content of water vapor contained within the working fluid.

Since the HPRTE water extraction capability so strongly motivated this experimental endeavor,

both the quantity of water extracted and the quantity not extracted were considered as key

performance parameters.

A problem presented by the gas analyzer was that it failed to include a direct quantification

of water vapor in the samples it measured. The analyzer only assessed the concentration of the

following species: 02, CO2, CO, NO, and a methane equivalent of UBHCs. A persisting

question was whether water vapor is considered by the machine when these concentrations are

output, or if they are dry concentrations. Unfortunately, this question was never satisfactorily

answered by the gas analyzer OEM, and some assumptions were necessary to complete the

analysis. It was assumed that the gas analyzer provided dry concentrations, according to

Equation 5.8,

nror = n + no2 +nco2 + nco + no +nUBHC (5.8)

where nTOT TepreSents the total number of moles being considered in a given sample. However










for this analysis, as it pertained to thermodynamic processes, the quantities of NO and UBHCs

were ignored, since they were several orders of magnitude smaller than the other constituents.

The CO mole fraction was also very small--less than 1%--but was retained for later service in

determining combustion efficiency. Given the absence of N2 COncentrations from the gas

analyzer output, Equation 5.9 was used to calculate the dry N2 mOle fraction, completing

knowledge of the dry exhaust gas composition.

YN2,DRY 01y2,DRY -CO2,DRY -CO,DRY (5.9)

Attention was then turned to the estimation of water content in each of the gas analyzer

samples, beginning with the assumption that the gas mixture was saturated upon reaching the

analyzer. Affirming this assumption was the consistent presence of condensate droplets inside a

transparent trap on the analyzer. Furthermore, this trap was consistently at or slightly above

room temperature. Combined with atmospheric pressure data, the partial pressure of water vapor

within the exhaust gases, PH20, WAS then estimated using saturated water properties from [15, pp.

924-925], providing the water mole fraction.

PH2
yH20 H2 (5.10)


Once again assuming the exhaust gases to behave as an ideal gas mixture, the summative partial

pressures of the dry constituents, PTOT,DRY, WAS determined with Equation 5.11i.

PTOT,DRY = P4,, PH20 (5.11)

Lastly, each of the dry mole fractions, yi, were corrected to include the water vapor, as shown in

Equation 5.12,










y, = 2,DRY TO,DRY (5.12)


where Pi,DRY TepreSents the partial pressure of the i.th dry exhaust gas constituent.

Basic Equations and Assumptions

The overall data reduction scheme was composed of a series of analyses performed on

each of the HPRTE components. In addition to the more obvious components such as

compressors and heat exchangers, various sections of ducting are also analyzed as

"components", since many of these have some kind of thermodynamic process occurring within

them. This section will provide a description of how each of the HPRTE components was

analyzed, including equations and assumptions utilized.

Main Air Inlet (MAI)

The MAI admitted air to the HPRTE during startup. It is an elliptical nozzle (or

bellmouth) with a 15.3 square-inch (98.7 cm2) throat area (AMAI). A pressure tap at the nozzle

throat was used to measure the differential pressure, APMAI, to atmosphere, and a J-type

thermocouple measured the temperature of the inlet air, TMAI. Atmospheric pressure and relative

humidity were also measured to ensure use of the proper gas constant, RMAI. While the MAI was

used only as a metering device for some of the HPRTE airflow, it was closed and admitted no air

for the maj ority of the Build 4 trials. Since the maximum Mach numbers at the throat were

around 0. 17 the air mass flow rate was arrived at using incompressible flow theory, and is shown

below in Equation (5.8).


tjhw;1 = Aw; ATMMM (5.8)
Rhw4 TMAI










Low Pressure Compressor (LPC)

The LPC also employed a bellmouth at its inlet, as well as a pressure tap connected to

instrumentation and a thermocouple at the inlet. The cross-sectional area of the LPC bellmouth

throat is 17.7 square-inches (1 14 cm2). The method of calculating the mass flow rate is shown in

Equation 5.9, and again derives from incompressible flow theory.


"iLPc = ALPCI PnALC (5.9)
RLPCI TLPCI

The LPC power (WLP ) WaS calculated using Equation 5.10:


WL~PC =LPc AhLPc (5.10)

where AhLPC WAS computed via the methods outlined in the Gas Properties subsection above.

The efficiency of the LPC was calculated from the definition of isentropic efficiency:

1'IPCl
PR /" -1
IlPC_ LPC (5.11)
LPCX. J
/LPCI

where PRLPC is the pressure ratio across the LPC, TLPCX TepreSents the LPC exit absolute

temperature, and TLPCI TepreSents the LPC inlet absolute temperature. The variable, YLPC,

denotes the average specific heat ratio across the LPC, per the Gas Properties section above.

Recirculation Venturi (RCV)

The RCV was used in the recirculation flow path exclusively as a metering device for the

recirculating gas flow. It was situated after the WGC in a straight run of 6-inch (15 cm) ducting

where the upstream pressure (PRCVI) and temperature (TRCVI) Were meaSured, as well as the

differential pressure (APRCV) fTOm upstream of the venturi to the throat. The throat area was

6.605 square-inches (42.61 cm2). Like the two inlet bellmouths on the HPRTE, principles from

incompressible flow theory were used to compute the velocity at the throat of the venturi, and in










turn, the mass flow rate. The equation used is shown in Equation 5.12,



n?2RCV = A (Rl PzAPC _A(.2
RRCV7RCly A,


where Al and A2 are the inlet and throat areas, respectively.

A primary operating parameter, the recirculation ratio, was then defined as the ratio of

recirculating gases to fresh air, as shown in Equation 5.13.


R = nR (5.13)
nLPC Af4I

Hot Gas Cooler (HGC)

The hot-side of the HGC was bounded by two, K-type thermocouples and had

approximately the same gas flow rate as that measured by the RCV. The difference in gas flow

rate is small, but worth noting in Equation 5.14 as follows:

liHGC RCV H20,iUC (.4

where viH2,,ifU accounts for the water extraction from the flow path before reaching the

recirculation venturi, but after the HGC. The heat load on the HGC hot side is computed as

shown below.


QHHGC Y2GC hHGC (5.15)

No attempt was made to compute the effectiveness of this heat exchanger, as insufficient data

existed on the cold side of the HGC--namely the cold side flow rate and fluid properties.

Warm Gas Cooler (WGC)

Both the hot and cold sides of the WGC were instrumented with thermocouples. The cold-

side utilized a paddlewheel flowmeter, and the hot-side flow rate was known to be the same as

that of the HGC. Accordingly, both the WGC gas-side heat rate (QifUC ) and water-side heat










rate (Qrectr ) could be calculated as follows:


QifUCG 2HGC 811, H 20fU heev (5. 16a)


Q rmer, = P wr ~if c p,W 9, T,: (5.16b)

In Equation 5.16b, pw is an assumed water density of 62.4 lbm/ft3 (1000 kg/m3), pi, is the

measured volumetric water flow rate in ft3/S, CPW is an assumed constant pressure specific heat

for water of 0.998 Btu/1bm-R (4187 J/kg-K), and ATw is the measured temperature rise of the

cooling water. In Equation 5.16a, hfg,WGC, iS the heat of vaporization. The heat of vaporization

was found using data adapted from [15, pp. 924-925] after computing the water partial pressure,

as shown below:

PH20 4H20 (PrmCI + Prizz,) (5.17)

where yH20 iS the mole fraction of water vapor in the gas stream, and PWGCI is the inlet pressure

of the WGC in units of psig. Again in Equation 5.16a, tiH2,,lfU denotes the quantity of water

condensed in and extracted from the WGC. Important to note at this point is that only the total

water extraction rate was directly measured--the summative water extracted from both the CGC

and the WGC. Because there were no flow measurements for the individual streams, a crude,

qualitative fraction was applied to the measured, total extraction flow rate, which estimated the

quantity of condensate issuing from each extraction point. For example, this fraction was

typically 60/40, for the WGC/CGC extraction rates.

It was most often the case that the two heat rates shown in Equations 5.16 were not

equivalent. The gas-side heat rate always tended to be higher than the water-side. This was

attributed to heat losses to ambient, which was straightforwardly found by adding the two heat

rates.









Efforts were made to calculate the effectiveness of the WGC. The water condensation was

one of the chief objectives motivating this experimental program. Thus, the proficiency of the

WGC to condense water was regarded as one of the HPRTE performance parameters, and was

accounted for in the effectiveness calculations, as detailed below. The effectiveness was

calculated as shown,


eWGC WG (5.18)
~maxWG'CC

where Qma,Wc is the maximum possible gas-side heat rate, as defined by Equation 5.19.

()max,WG;C (IjZWGCNrC AhmxSVN 72max,WGC,H20(Amax,SEV~,H20 +hfgWC (5.19)


In Equation 5.19, tizWG,Nc is the non-condensable gas flow rate, Ahmax,SEN,NC is the change in

enthalpy if the non-condensable gases are cooled to the water-side inlet temperature, and

ri2ma,WG,H20, iS the estimated, maximum possible water extraction rate if the mixture is cooled to

the cold-side inlet temperature. The variable Ahmax,SEN,H20 TepreSents the sensible change in

enthalpy from cooling the water vapor to its initial dew-point. Again, this treatment effectively

considers the flow as two separate streams--one condensable stream and one non-condensable

stream--and rests on the assumption of an ideal gas mixture, even for the water near its vapor

dome. Furthermore, it is understood that the interaction of the water vapor and the non-

condensable stream will produce effects not captured by this analysis. Specifically, the water

vapor will experience further sensible enthalpy changes below the initial dew-point, since the

dew-point temperature will gradually decrease as more water is taken from the mixture.

Cold Gas Cooler (CGC)

A mixing junction was situated directly upstream of the CGC where fresh air from the LPC

joined with recirculating exhaust gases. Mixing junction calculations were necessary to find the










mixed, inlet temperature to the CGC, and are discussed in the Mixing Junction section later. The

CGC exit temperature was acquired with a J-type thermocouple. The CGC heat rate was

estimated with effectively the same equation as that used for the WGC, as shown in Equation

5.20:


QCGC R jZCV aCGC +jH20,CGChfg,CGC (5.20)

where viH2,CGC,, TepfeSents the mass flow rate of water extracted from the CGC. Again in

Equation 5.20 hfg appeafS, but has been recalculated at this point for both the reduced mole

fraction of water vapor and gas pressure at this state point.

The effectiveness calculations performed for this heat exchanger were also identical to

those of the WGC, shown above in Equations 5.18 and 5.19.

Ducting Section 4-5

This ducting component was defined by the ducting between the HPC exit thermocouple

and High Pressure Recuperator (HPR) inlet thermocouples. A heat balance was deemed

necessary for this ducting section due to the placement of the HPC exit thermocouple. The

placement of this thermocouple by the OEM resulted in the measurement of HPC exit gases in

the middle of an incidental heating process. Upon emerging from the diffuser, the HPC exit

gases passed through an annular passage over hot, HPT inlet ducting before its egression from

the engine. The result of this was significant heating between the two thermocouples defining

this ducting section. This heating rate was computing using Equation 5.21:

Q4-5 =jHPC h4-5 (5.21)

where viHPC WaS calculated by applying the conservation of mass principle to the HPRTE, per

Equation 5.22.

#HPC AMI1 LPC RCV H 20,CGC (.2










High Pressure Compressor (HPC)

As mentioned above, there was significant heating of the gases leaving the HPC. One of

the initial obstacles in characterizing the HPC performance was to determine the HPC adiabatic

exit temperature. With the HPC exit thermocouple residing in the middle of this inadvertent

heating, it was necessary to utilize some of the engine manufacture data on the HPC. Included

within the engine documentation [16] was a uniquely formatted compressor map for the HPC

that served this purpose. This map is shown in Figure 5-1, and provides the adiabatic HPC exit

temperature as a function of corrected flow rate and corrected speed. The HPC inlet and

adiabatic exit temperatures were then used to find its Ah, and in turn its work rate, as shown

below.

WePC = tjHPe~hHPe (5.23)

The HPC isentropic efficiency, TIHPC, WAS computed with the same equation aS TILPC, Shown in

Equation 5.11.

Similar to ducting section 4-5, a heat rate was defined to account for the temperature rise

between the one-dimensionalized, adiabatic HPC exit and the HPC exit thermocouple. This heat

rate, shown in Equation 5.24 was arbitrarily grouped among other HPC parameters and

considered as a heating process immediately subsequent to compression, in order to maintain

continuity of the flow path. The enthalpy change in this heating process (Ah4) WAS computed

using the adiabatic HPC exit temperature found via Figure 5-1, and the temperature acquired by

the HPC exit thermocouple.

Q4 = jHPC h4 (5.24)









Combustor

The combustor presented several challenges during the formation of this data reduction

algorithm. Principle among these was the less-than-ideal location of the combustor exit

thermocouple. This thermocouple was situated about 14 inches (36 cm) from the flame, a

radiation source approaching 3500 OF (2200 K), and only partway through the mixing of hot

primary-zone gases and cool dilution gases. Moreover, for reasons discussed later, the HPT exit

temperature was also difficult to ascertain, making it difficult and unreliable to back calculate the

HPT inlet and combustor exit temperatures with a work balance. Thus, a different approach was

adopted to estimate the combustor exit temperature as accurately as possible using the fuel flow

rate, and combustor inlet conditions. Initial efforts approximated the combustor as a heater with

no losses to ambient, and later evolved to estimate convective and radiative losses to ambient.

The rate of heat added to the working fluid in the combustor was estimated using Equation 5.25:

Ocom = (ilFELLH)9cam -Ocor R4D(5.25)

where LHV is the fuel lower heating value, Qcour and QR4D are the estimated convective and

radiative rates of heat loss from the combustor to ambient. The type of fuel used for these

experiments was diesel-2, and its LHV was assumed to be 18,300 Btu/1bm (42.6 MJ/kg). The

combustion efficiency, TICOMB, WAS defined using the CO and CO2 mOle fractions (yco and yCO2)

obtained by the gas analyzer, and rests on the assumption that the exhaust gas composition was

frozen between the combustor exit and the exhaust stack where gas samples were measured.


reoaM = 1 _7 (5.26)


To estimate the two heat losses in Equation 5.25 the combustor was discretized into several

sections defined by its geometric features. During engine runs, a surface probe was used to









obtain surface temperatures on the combustor. Equations 5.27 show how each of the heat losses

were computed:

Ocour = hC Az (TSRF~ -TM) (5.27a)


()RA = Ecrf Az (TS4URF,z T ~) (5.27b)


where in each case the losses from each discrete area (Ai) and its corresponding surface

temperature (TSURF,i) are Summed over the surface of the combustor. TAMlB TepreSents both the

temperature of the ambient air and that of the radiative heat sink, and a represents the Stefan-

Boltzmann constant. In Equation 5.27b, the radiative heat exchange is approximated as that

between two black bodies, and as such the emissivity, E, is one. In Equation 5.27a, h is an

assumed convection coefficient of about 25 W/m2-K. Incropera and DeWitt [15, p. 8] suggest

that this value is at the high end of the natural convection regime, and at the low end of the

forced convection regime. The actual heat transfer environment around the combustor had a

very low velocity flow passing over it, induced by the facility fans, and as such was judged as

forced-free convection. In any case, the estimations obtained from Equations 5.27 were

consistently less than 1% of QCOMB

Once the heat addition rate was computed from Equation 5.25, the enthalpy change of the

working fluid in the combustor, AhCOMB, WAS computed with Equation 5.28:


Ah .=cos(5.28)


where vcous is simply the sum of viHPC and MFUEL From this point, given the combustor inlet

temperature, a subroutine was employed to iterate on the exit temperature until Ah equaled

AhCOMB. The iterative process was made necessary by the quartic polynomials, in temperature,

that were used to compute Ah (Equation 5.5).










The equivalence ratio (0) was calculated using Equation 5.29:


FA
O = (5.29)
FSTOICH

where FA is the actual fuel-to-air ratio, and FASTOICH is the stoichiometric fuel-to-air ratio. The

fuel-to-air ratio was computed with Equation 5.30.


FA = nUE (5.30)
namz +nLP

It is important to note, as pointed out by Crittenden [17], that the calculations shown in

Equations 5.29 and 5.30 fail to consider unburned oxygen contributions from the recirculation

gases. Therefore, the combustion environment would actually be leaner than Equation 5.29

would imply, if there is recirculation flow.

Ducting Section 7-8

This section of ducting exists between the combustor exit and the HPT inlet, and is

characterized by its heat loss to gases exiting the HPC (Ducting Section 4-5). In a sense, these

two segments of the HPRTE flow path can be viewed as an internal heat exchanger. The

assumption is made both here and for Ducting Section 4-5 that this hypothetical heat exchanger

is well insulated from ambient, since it is the cold-side that shares a surface with ambient air.

Proceeding from this assumption, the heat lost by ducting section 7-8 is equated to the

cumulative heat gained by the HPC exit gases before reaching the recuperator, per Equation

5.31.




The next step was to use this heat rate to find the enthalpy change, Ah7-8,


Ah,_, = '(5.32)
nCOMs









and iterate to find its corresponding HPT inlet temperature, with the previously calculated

combustor exit temperature.

High Pressure Turbine (HPT)

As was previously mentioned, there were significant uncertainties in directly acquiring the

HPT exit temperature. The OEM equipped the HPT exit with four compression fittings to

accommodate four HPT exit thermocouples. These are situated around the circumference of the

exhaust plane, 90 degrees apart. Specifically, the uncertainties arose from the apparent, highly

stratified flow exiting the HPT. Each of the four thermocouples consistently indicated four,

distinctly different temperatures differing by as much as 320 OF (178 K). Rotation of the

thermocouples from one location to another confirmed that the phenomenon was particular to the

temperature field around the HPT exit plane, and not the thermocouples themselves. It was

believed that simply averaging the four temperatures would be a poor representation of the

mixed, one-dimensional HPT exit temperature, since initial efforts to use a linear average

resulted in erroneously high isentropic efficiencies (TIHPT > 100%). Consequently, a different

approach was employed, seeing as enough information was available from elsewhere to

complete the HPT analysis.

Instead of attempting to directly calculate the HPT power by means of the fluid Ah, a

power balance was implemented, as shown in Equation 5.33:

HPr WnC DYNO -Woss (5.33)

where WDYN is the power absorbed and measured by the dynamometer, and Wross represents

the power consumed by friction and pumps within the engine. The power losses were provided

by the OEM [16] for each of components internal to the core engine: fuel pump, oil pump,

bearing friction, and rotor friction. In addition, the total power loss data was conveniently added










by the manufacturer. All of these losses are shown as a function of rotor speed, and can be seen

in Figure 5-2. A built-in Matlab function was utilized to interpolate from the total power loss

curve using a cubic spline interpolation method.

The dynamometer power was calculated using measurements acquired from a load cell that

countered the engine torque. For convenience, a straightforward equation was provided [16] by

the dynamometer manufacturer, as shown below:


WD1NO FD1NO D1NO (5.34)
4500

where FDYNo and 10DYNo are the force (in pounds) acquired by the load cell, and dynamometer

speed (in rpm), respectively. The number in the denominator is the dynamometer constant, and

accounts for the length of the dynamometer moment arm, and also for unit conversions such that

WD1N is in units of horsepower.

Having found the HPT power, and with knowledge of the HPT flow rate, the enthalpy

change across the HPT (AhHPT) COuld then be computed with Equation 5.35.


Ah,, = "'p (5.35)
ncohlm

Again with the component inlet temperature as a starting point, an iterative subroutine was used

with AhHPT to find the HPT exit temperature. Upon calculation of the HPT exit temperature, the

HPT isentropic efficiency was computed using Equation 5.36:

1-HTX
/T
rler = anPT (5.36)
1 PRHP HP

where THPTI and THPTX are the HPT inlet and exit absolute temperatures, respectively. The









variable, YHPT TepreSents the average specific heat ratio across the HPT, and PRHPT is the HPT

pressure ratio defined in Equation 5.37.


PRHPT rPHTX +PAMr (5.37)
PHPTI + PATM

Recuperator

For most of the Build-4 engine runs the recuperator was well insulated with a two-inch (5

cm) thick, ceramic blanket on top of another thin, fiberglass wrap. This effort was primarily

made to simplify the data reduction algorithm, reducing or eliminating the diffacult-to-ascertain

heat loss to ambient from the recuperator. The surface temperature of the insulation was

measured during engine runs, and a rudimentary analysis confirmed that losses to ambient were

typically three to four percent of the heat transfer rate between the hot and cold streams.

Nonetheless, the estimation of heat losses through the insulation was retained in the analysis.

For engine runs with the insulation in place, the hot-side heat rate calculation is shown in

Equation 5.38 using the cold-side heat rate and the abovementioned heat rate to ambient:

QRHOT RCi~OLD RAMB_~L (5.38)

where QR),AM TepfeSents the heat losses to ambient, and QR,COLD is the recuperator cold-side heat

rate, computed with Equation 5.39,

QR,COLD = jHPC hR,COLD (5.39)

where AhR,COLD WAS calculated with the two well-placed thermocouples bounding the cold-side

of the recuperator. Similar to the combustor heat losses to ambient, the recuperator QR,AM WaS

calculated by breaking the recuperator into three discrete areas and surface temperatures, and

then applying Equation 5.40.










GR,AMB~ = -hf AzAT (5.40)


The enthalpy change across the recuperator hot side was then calculated using Equation 5.41,

OR,HOT
AhRO (5.41)


which in turn enabled the recuperator hot-side inlet temperature (also the HPT exit temperature)

to be re-calculated using the recuperator hot-side exit temperature. This new HPT exit

temperature was compared to the previously calculated value to make sure they corresponded

satisfactorily. From this point, the HPT exit temperature was used with AhHPT to again make

redundant calculations of HPT inlet and combustor exit temperatures and ensure agreement with

the previously calculated values.

For engine runs that took place before the insulation was utilized a slightly different

method was employed. Due primarily to the large variations in the recuperator surface

temperature, the heat rate to ambient was estimated with much less confidence. Additionally,

this heat loss was considerably larger, and as such its uncertainty would weigh more heavily on

the uncertainty of the hot-side heat rate. For this reason, redundant calculations were not made

in these cases, and the recuperator hot-side heat rate was computed with Equation 5.42,

QR,HOT = i)Th (5.42)

where in this case AhR,HOT WAS found using the previously calculated HPT exit temperature and

the measured hot-side recuperator exit temperature. The ambient heat losses for the uninsulated

cases were found by summing the hot-side and cold-side heat rates.

Lastly, the recuperator effectiveness was computed using Equation 5.43.


E = (5.43)
GR,MX










Due to its lower average temperature and specific heat, the cold stream consistently had the

lower heat capacity rate. Therefore, the cold-side stream is the only one that can achieve the

maximum Ah, and the maximum heat transfer rate is defined in Equation 5.44,

QR,Af4X YiHPC Ah (5.44)

where AhR,MAX was found using the recuperator cold-side and hot-side inlet temperatures.

Low Pressure Turbine (LPT)

No flow rate measurements were practical given the ducting arrangement of the LPT, only

pressures and temperatures at the LPT inlet and exit were acquired. However, a power balance

with the LPC made a LPT flow rate estimate possible. This was accomplished by assuming a

mechanical energy conversion efficiency, TlhfECH, Of 0.97 between the two, allowing the LPT

work rate to first be estimated, as shown in Equation 5.45.

-W
WLPT __LPC (5.45)
rhfECH

After calculation of the enthalpy change across the LPT (AhLPT) the LPT flow rate was estimated

using Equation 5.46.


n?2LPT _LPT (5.46)
hLPT

A LPT map plotting corrected flow rate over LPT pressure ratio was also available to provide a

redundant flow rate estimate. The two values consistently agreed satisfactorily with one another.

The LPT isentropic efficiency was calculated using Equation 5.36 with LPT pressures and

temperatures substituted.

Mixing Junctions

There were several mixing junctions in the HPRTE flow path. Only one of these,

however, required an energy balance to find the mixed exit temperature. This mixing junction









existed just upstream of the CGC, where the LPC exit gases mix with recirculation gases issuing

from the WGC. In this case, the mixing process was treated as adiabatic. Furthermore, the

temperatures and specific heats of the two inlet streams were close enough to one another to

justify the approximation of constant specific heats. Used here only, this approximation

permitted the temperature of the exit stream to be straightforwardly found using Equation 5.47,

which presents the calculation using generalized variables.


L3 2(dzcP, + 2 PC,, Cn (5.47)


In Equation 5.47, the subscript, 3, denotes a variable pertaining to the mixed, exit stream, and the

subscripts 1 and 2 denote variables pertaining to the two inlet streams. The specific heat of the

mixed, exit stream is approximated as the linear average of the two inlet streams' specific heats.

Changes in gas composition were also tracked across mixing junctions. It should be

pointed out that water extraction locations were also classified as mixing junctions. Although

there was no mixing taking place, per se, a sort of "un-mixing" of fluids was occurring.

Inclusion of water extraction locations with the more typical mixing junctions allowed use of the

same equations, generalized for gas composition bookkeeping. The mass flow rate leaving a

mixing junction was plainly found with Equation 5.48:




where the sign of the second mass flow rate is determined by whether the fluid stream is entering

or exiting the mixing junction. Where in Equation 5.47 the subscripts 1 and 2 denoted the two

inlet streams, they shall henceforth denote the two Imown streams, since information about these

two streams is known, and information pertaining to stream 3 is unknown. Next, the molecular

weights of the two known streams (M1,2) WeTO COmputed using Equation 5.2b and the mole flow

rates of all three were found using Equations 5.49:













1,= 2n


(5.49a)


(5.49b)


where again, in Equation 5.49b, the sign of the second known stream depends whether it' s

mixing with stream 1, or exiting the flow path. From here, the mole flow rates of each stream

were multiplied with their corresponding vector of mole fractions to obtain mole flow rates of

each pure substance within each of the known streams, as shown below.


nN2 N2
no 2 O 2
nCO 2 i1,2 yCO 2(5
nH 20 H 20
iCO _I 1, 2 O 2

Finally, the mole flow rates from the known streams were used to find the mole fraction vector

of the unknown stream, as seen in Equation 5.51i.


5


0)


yN2 N2, N2,
y02 02, 02
1 1
yCO2 CO2 CO (5.51)

H20, H20 H20
yco o 3 O_1 _CO_2

In the special case of water extraction locations, the sign in front of stream 2 was negative, and

the mole fraction vector elements of stream 2 were all zero, except for the fourth element, which

was one.

Propagation of Uncertainty

In addition to reducing and interpreting the experimental data, a thorough propagation of

uncertainty analysis was performed. To complete the uncertainty analysis, classical methods

were employed as presented by Holman [18]. For a given dependent parameter, R, with n










independent parameters as shown in Equation 5.52,

R = f(xl,,x ,..., x,,) (5.52)

the uncertainty of R was computed using Equation 5.53,



mRI X1X "X (5.53)


where coxn is the uncertainty of the nth independent parameter. Equation 5.53 was applied to find

the uncertainty of all the parameters discussed in the above section.

All uncertainties of reduced parameters stemmed from one of three possible sources:

instrumentation error, data acquisition error, and miscellaneous transient behavior, either real or

artifact. This concept is generalized by Equation 5.54.

Ox~ = Oxasr~ + 0x,moQ + 0xxxas (5.54)

Ideally, all three sources would be comparable and small. However, it is believed that the

three contributions usually differed considerably, and were most frequently dominated by the

third term. Typically, data were collected over a period ranging from five minutes to 15 minutes,

and during this time nearly all of the temperatures, pressures, and flow rates either climbed, fell,

or fluctuated to some degree. Since the obj ective was to characterize a single, steady operating

point by averaging several hundred data points, many of the uncertainties were often taken as

one standard deviation from the mean. For some parameters, such as ambient temperature, there

was very little fluctuation aside from some noise in the signal. In such cases, the uncertainty in

the instrument itself was the dominating source of error.

Error derived from the AD conversion of signals was considerably smaller than the other

two sources. 16-bit resolution was used within the DAQ system to curb quantization errors, and

the time between samples was sufficiently large to avoid any appreciable aperture errors.










Occasional testing was also carried out to spot-check for significant non-linearity errors made by

the AD conversion.

In general the uncertainty of all the measured values was taken to be the greater of the

instrument error and one standard deviation from the mean value. Only one standard deviation

was thought necessary because the data were typically skewed by relatively few outliers. These

outliers could result from several occurrences, e.g. the momentary loss of a thermocouple signal,

or indeterminate electromagnetic interference. The dynamometer water supply pressure would

also fluctuate erratically due to other plumbing loads in the building. This often caused brief

excursions in the engine load, engine speed, fuel flow rate, and other parameters as well.

Instrument uncertainties for thermocouples, pressure transducers, and other instruments

can be found in Table 5-1.

Data Adjustments

The reduced data described in the above sections were obtained from a very particular

engine and combination of components. Consequently, some of the capabilities of this combined

cycle were veiled by excessive pressure drops, poor VARS performance, and high ambient/PCW

temperatures. In addition, comparisons between these experimental results and other

experiments or models needed to be made on the same basis. For these reasons, classical,

partially non-dimensional parameters were used to present the data in a properly scaled form.

The corrected parameters chosen are explained in greater detail by Volponi [19], and are shown

in Table 5-2.

As mentioned above, the pressure drops in the HPRTE ducting was high enough to obscure

the experimental results. Specifically, when transitioning into the pressurized regime, the HPT

was initially back-pressured in order to accelerate the LPT. This resulted in a premature increase

in the HPT exit and inlet temperatures, such that very little or no margin was left for loading the










engine by the time the LPC exit pressure rose appreciably. Furthermore, various leaks in the

ducting and engine casing intensified with increasing boost due to the increased pressure within,

resulting in an un-metered loss of working fluid. Consequently, nearly all of the pressurized data

sets were at or near maximum HPT inlet temperatures, and at reduced engine speeds with no

load on the dynamometer, resulting in an uninformative scatter of net power around zero. This

necessitated a supplementary modeling approach to approximate the net power if the pressure

drops had been different, since in reality this can be an arbitrary design choice.

Pressure Drop Considerations

In the interest of better evaluating the HPRTE performance, several adjustments were

made to the raw data, and a supplementary modeling approach was employed. In this modeling

approach, the empirical pressure drops in the recirculation flow path and within the recuperated

engine. This provided an upper-bound estimate of HPRTE power and efficiency in the absence

of any pressure drops. The first pressure adjustment reduced the hot-side recuperator and

recirculation pressure drops to zero by setting the HPT exit pressure equal to the HPC inlet

pressure. The second adjustment eliminated the core-engine pressure drop by raising the HPT

inlet pressure to the HPC exit pressure.

The above adjustments translate to increased net power in that the HPT pressure ratio is

increased, resulting in a greater temperature ratio and temperature difference. The new HPT

expansions began from the same temperatures, but were completed at lower temperatures. The

lower HPT exit temperatures were calculated by assuming the HPT efficiency to be constant in

all cases. Though no maps were available for this turbine, this was judged as a fair

approximation when considering a map generated by AXOD [20] in Figure 5-3. There it can be

seen that the HPT efficiency is approximately 85% for a wide range of corrected speeds and

pressure ratios bounding the range of values seen in experiments. The adjusted HPT exit










temperature was computed using Equation 5.55,


THPTX THPI HPTpT HP" 5.55



where TIHPT is 0.85. Having estimated the adjusted HPT exit temperature, the adjusted net work

output from the HPRTE was computed using Equation 5.56.

WNET = k h -HPC (.6

In the interest of further generalizing the reduced results, it can also be seen from Equation 5.56

that the friction work has been ignored in this computation. The HPC work term, WHPC, WaS

calculated using Equation 5.23.

Table 5-1. Uncertainty of instruments
Instrument Uncertainty

T-Type Thermocouples Greatest of 0.9 OF (0.5 K) or 0.4%

J-Type Thermocouples Greatest of 2.0 OF (1.1 K) or 0.4%

K-Type Thermocouples Greatest of 2.0 OF (1.1 K)or 0.4%

0-30 Pressure Transducers 0.1% of F.S. (typical), 0.5% (max)

0-100 Pressure Transducers 0.1% of F.S. (typical), 0.5% (max)

Differential Pressure Transducers 0.1% of F.S. (typical), 0.5% (max)

Chilled Water Flowmeters 0.5%

Fuel Flowmeter 1.0%

Water Extraction Load Cell 0.2 % of F.S.

Dynamometer Load Cell Unknown

Engine Speed Optical Tachometer 3 rpm










Instrument


Uncertainty

UBHC: 20 ppm
CO: 0.03%
CO2: 0.6%
02: 0.2%
NO: 10 ppm


Gas Analyzer


Table 5-2. Data scaling parameters [19].
Parameter


Corrected Parameter
N
N* =





Wi* =

**_T


Rotor Speed


Gas Mass Flow Rate


Power


Temperature

Pressure


P*= P




























































Figure 5-1. Map used for finding the HPC adiabatic exit temperature, shown as a function of
corrected flow rate and corrected speed [16].



76














































































|20,000 30,000 ENG1NE SPEED 48000 50000


Figure 5-2. Rover 1S-60 internal power losses as a function of rotor speed [16].


bnrnn












Rover Turbine Efficiency Map


1 .5 2 2.5 3 3.5
Pressure Ratio


Figure 5-3. Rover 1S-60 turbine map generated by AXOD [20].









CHAPTER 6
EXPERIMENTAL RESULTS

The experimental apparatus described in Chapter 3 was operated according to the

procedures discussed in Chapter 4 on 13 occasions to yield 34 distinct sets of working data. The

data reduction algorithm outlined in Chapter 5 was then applied to the data, providing the

relevant performance parameters. As was mentioned in Chapter 5, all but five of these data sets

were idle runs, this was due primarily to excessive pressure drops and a slightly mismatched

turbocharger. Since no load was applied to the dynamometer, the resulting data--specifically

the net power and efficiency--were less informative, as shown in Figure 6-1. Since the engine

was in a sense loaded by back-pressuring the HPT, the steps outlined at the end of Chapter 5

were taken to estimate the power consumed by the various loss mechanisms and present the

adjusted power. All of the adjusted, reduced data generated by the methods discussed in Chapter

5 are tabulated in Appendix C. Worth noting is that all of the data sets were at or near the

maximum sustainable HPT inlet temperature, and as such represent full power engine runs, while

other operating parameters such as recirculation ratio and LPC pressure ratio were varied. For

conciseness the values previously referred to as adjusted power will henceforth be referred to as

power, since it is the only power of interest.

Gas turbine engines exhibit a particularly strong dependence on ambient conditions,

namely pressure, temperature, and humidity. However the HPRTE is unique in that the inlet

pressure of the core engine can be changed by controlling the LPC pressure ratio. Furthermore,

when coupled with the VARS, the HPC inlet temperature can also, in principle, be changed at

the will of designers and operators. For these reasons, 6 and 6, the classical correction

parameters defined in Table 5-2, were calculated using the HPC inlet conditions instead of the









ambient conditions, since ambient conditions possess little influence over the HPRTE

performance.

Although most turbine engines, including the Rover 1S-60, have the capability of

operating at various speeds, no attempts were made in these experiments to capture data at off-

design speeds. However, as the HPT was back-pressured and the turbocharger accelerated, there

was significant speed-droop resulting from the mechanical governor, which helps to validate the

idea that the engine was, in effect, being loaded. Additionally, the corrected speed (N*), defined

in Table 5-2 showed even greater variance, since the HPC inlet temperature fluctuated as well.

Since N* refers performance parameters to standard ambient conditions, it was used as a

dependent parameter in place of engine speed (N). In the interest of characterizing the

performance of this engine, several plots were generated. The first of these is Figure 6-2, which

plots corrected power (P*) and thermal efficiency against N*. Also included within Figure 6-2 is

the adjusted power before inlet-condition corrections (P). A peculiar characteristic of the P and

P* data is that they have opposite slopes when plotted over N*. This results from the fact that

the scaling parameters generally demonstrated a dependence on engine speed. Ideally, any range

of HPC inlet pressures and temperatures could be actualized at any operating point (any N* for

example), but due to the operational boundaries of the engine and VARS, these inlet conditions

were coupled with engine speed.

The scaling parameters, 6 and 6, are shown over N* in Figure 6-3 for comparison with

Figure 6-2. I can be seen in Figure 6-3 that both scaling parameters increase when decreasing

N*. In the case of 6, this trend is more straightforward. The higher LPC pressure ratios

(essentially the same values as 6) were attained only at lower engine speeds, because increased

operation of the turbocharger directly caused the speed-droop, as described in the Data










Adjustments section of Chapter 5. The 6 trend also appears to follow speed, but more directly

correlates with LPC pressure ratio, with some random fluctuations as well. Again, as the

turbocharger was accelerated, greater flow rates were diverted from the exhaust and driven

through the LPT and through the recirculation flow path. In addition, as the HPT back-pressure

initially rose, the HPT pressure ratio decreased until the LPC could be accelerated enough to

compensate, increasing the HPC inlet pressure. This caused the exhaust gas temperature (EGT)

to climb. As the engine speed dropped, the HPT pressure ratio continued to fall as well, further

increasing EGT. These two consequences of pressurization-increased recirculation flow and

temperature-combined with one another to greatly increase the heat load on the VARS

generator (the HGC) in particular, but on the WGC and CGC also. The resulting heat loads

taxed the VARS and the WGC exceedingly and always effected an increase in HPC inlet

temperatures. Thus, 6 tended to scale with LPC pressure ratio, and in turn speed. The

relationship between 6 and LPC pressure ratio is shown in Figure 6-4. When considering the

trends in Figure 6-3, it is important to point out that 6 itself appears in the N* parameter as

shown in Table 5-2. As a result, the obvious decrease in N* when decreasing N is compounded

by the related upsurge in HPC inlet temperature.

After considering the trends of the correction parameters relative to N*, the opposite slopes

of P and P* can be better understood. Since both 6 and 6 appear in the denominator of P*, and

both have downward slopes over N*, it is more reasonable to expect the results shown in Figure

6-2 provided the denominator of P* (31 ) changes faster than its numerator (P). In general, P

would be expected to scale one-to-one with HPC inlet pressure such that a given increase in 6

would be accompanied by an equivalent increase in P. However, competing effects somewhat










mitigated the desired power-pressure relationship. Power is plotted against LPC pressure ratio in

Figure 6-5.

Figure 6-5 shows the power obtained from the adjusted data, as well as the power

predictions of a model developed by Khan [8, 12]. The model is a steady-state, one-dimensional

thermodynamic model of the HPRTEn/VARS. The adjusted experimental power and modeled

power differ significantly, as seen in Figure 6-5. An initial, sharp rise in P is observed over LPC

pressure ratio, but a weaker dependence at higher pressure ratios. The modeled results, however,

predict a steady, linear increase in power over LPC pressure ratio. The magnitude of power

predicted by the model is most likely lower due to the inclusion of friction, parasitic horsepower,

and pressure drops by Khan, in an effort to most accurately represent this engine and its

configuration. The adjusted power, shown in blue, was an upper-bound estimate using neither

pressure drops nor friction power. The anticipated rise in P is more directly related to the engine

mass flow rate, which itself depends on the HPC inlet pressure assuming the inlet temperature is

held fast by the VARS. In actuality the inlet temperatures crept up with the higher LPC pressure

ratios (Figure 6-4), lessening the density and potential mass flow rate, although this was a weak

effect. A more prevailing detraction from the potential flow rate was the speed-droop also

associated with higher inlet pressures (Figure 6-3). The deceleration of both the HPC and HPT

caused a diminution in the flow rate through, and pressure ratio across each. With the reduced

pressure ratios came also the reduction of temperature ratios, and therefore smaller temperature

differences across the high pressure turbomachinery. In consequence, while the mass flow rate

did manage to rise alongside of the LPC pressure ratio, the HPC and HPT temperature

differences dropped. Figure 6-6 illustrates these opposing trends in a dimensional sense, while

Figure 6-7 shows the same data, but normalized. The mass flow rate data in Figure 6-7 are









normalized by their minimum, and the delta-T data by their maximum. In this way the percent

changes in each of these parameters can be viewed, and the slopes compared. It can be seen in

Figure 6-7 that despite the abovementioned mitigating factors, the percent increase in mass flow

rate was greater than the decline in turbomachinery temperature differences. The manner in

which these parameters influence power can be seen in Equations 5.56 and 5.23. Another

suspicion that can be verified in Figure 6-7 is that the mass flow rate increase was less than

expected. This is realized by noting that the slope of percent-change-in-flow-rate regression line

indicates only an 84% increase in flow rate, per unit increase in LPC pressure ratio. Again, with

ideal hardware and control schemes, an approximate 100% increase in flow per unit LPC

pressure ratio increase is expected.

The engine mass flow rate and corrected mass flow rate are shown over N* in Figure 6-8

for comparison with Figure 6-2. The trends in Figure 6-8 are similar to those in Figure 6-2 in

that the uncorrected values bear a downward slope and the corrected values an upward slope.

This comparable behavior can in like fashion be explained by the relationship 6 and 6 exhibit

with N*. One difference to be noted is the relatively gentle slope of corrected flow rate over N*,

where P* demonstrated a much closer relationship with N*. However, 60.5 appears in the

numerator of the corrected flow rate parameter, and in the denominator of the corrected power

parameter. Thus the downward trend of 6 with increasing N* serves to somewhat lessen the

impact of 8 instead of augmenting it, as was so for P*. Thus a greater mass flow rate (and 6) at

lower speeds was more than balanced by the greater HPC inlet pressure.

Having considered the dependence upon inlet conditions and mass flow rate, other non-

dimensional parameters and their effects on power were also explored. In particular, two cycle

parameters were thought to have a significant bearing on power: the HPC pressure ratio, and the









overall Cycle Temperature Ratio (CTR). The CTR is defined as the ratio of the maximum,

absolute cycle temperature (HPT inlet temperature) to minimum absolute cycle temperature

(HPC inlet). These parameters' influence can most easily be understood by examining a

temperature-entropy (T-S) diagram with an ideal Brayton cycle, shown in Figure 6-9.

The diagram in Figure 6-9 shows all temperatures normalized to the inlet temperature such

that CTR appears on the ordinate. Three different cycles are shown on the T-S plot for

comparison: a baseline cycle, a higher CTR cycle at the same pressure ratio as baseline, and a

higher HPC pressure ratio at the same CTR as baseline. The baseline cycle follows the path 1-2-

3-4, and the power-per-unit-mass is equivalent to the area contained by the cycle, areas A plus B.

The second cycle follows the path 1-2-5-6, and clearly yields a significantly higher specific

power with areas A, B, and C. The third cycle follows the path 1-7-8-9. This cycle differs in

that it adds to the specific power with area D, but subtracts from specific power by surrendering

area B. In general, there will be a net gain in specific power, but a cycle having different

proportions when sketched on a T-S diagram could conceivably break even or lose specific

power. For this reason, specific power is expected to scale more strongly with CTR than with

HPC pressure ratio, as is made evident by the data shown in Figure 6-10.

Figure 6-10 shows specific power over CTR, but the data are grouped according to their

HPC pressure ratios so that the influence of each cycle parameter can be examined

independently. Regression lines were added only to better illustrate trends made less distinct by

the data-scatter. Several of the lines are fit to only three to four data points, and their slopes can

be misleading. The data in Figure 6-10 clearly show the closer relationship between CTR and

specific power. Additionally, at a constant CTR, specific power climbs somewhat with HPC

pressure ratio, but to a lesser, even diminishing extent at higher pressure ratios.









In Figure 6-11, P* is also presented over CTR for various HPC pressure ratios. The data in

Figure 6-11 indicate a somewhat stronger influence of HPC pressure ratio on P* than of CTR on

P*. This is most likely related to the closer dependence of HPC pressure ratio on N*. HPC

pressure ratio followed N* very closely, where CTR had much less dependence thereupon, as

shown by Figure 6-12. Though, incidentally, the HPC inlet temperature (the denominator in

CTR) did tend to follow speed, that relationship is not as strongly coupled as that of HPC

pressure ratio and N*. Furthermore, the HPT inlet temperature had no reason to trend with speed

since it is an independently controlled variable, and all of these data sets were maintained at or

near the engine maximum HPT inlet temperature.

The cycle thermal efficiency may also be expected to behave similarly relative to CTR and

HPC pressure ratio. The efficiency data exhibit more scatter than the power data, and for this

reason two separate Eigures are used to show their dependence on CTR and HPC pressure ratio.

Figure 6-13 shows efficiency over CTR and Figure 6-14 plots it over HPC pressure ratio. In

both charts the thermal efficiency demonstrates to a small extent the anticipated, upward trend

with CTR and HPC pressure ratio. Though these data are more widely dispersed, the stronger

correlation with CTR can again be discerned. It is more likely that the ostensible, direct

relationship between HPC pressure ratio and thermal efficiency is artifact, since the higher

pressure ratios happened to coincide with the higher CTRs. More importantly, the higher

pressure ratios also occurred together with higher corrected speeds. Figure 6-15 shows that the

HPC isentropic efficiency increased with N* as it approached the design speed of 46,000 rpm.

The more efficient HPC operation probably contributed the most to any apparent relationship

between overall thermal efficiency and HPC pressure ratio. Any similar effects from the HPT

efficiency are lost, since it was Eixed at 0.85 for the power adjustment calculations. Overall,










there was little variation in thermal efficiency, largely because all of the data points are at full-

power.

A principal dimensionless parameter associated with the HPRTE is the recirculation ratio

(R) defined by Equation 5.13. This parameter indicates the extent to which the engine is

operating within the semi-closed regime. It is generally controlled by changing the LPC pressure

ratio, but also depends on the differential pressure between the hot-side recuperator exit and HPC

inlet state points. In a physical sense, this is the pressure drop of the recirculation circuit. The

Build 4 engine configuration was equipped with a valve in the recirculation flow path with which

to vary this differential pressure. This valve, shown as VREC in Figure 1-1, proved essential in

making transitions into the pressurized regime and providing a Einer control over recirculation

flow rate, and over R.

Figure 6-16 shows the recirculation ratios against LPC pressure ratio. Again, the modeled

results are included for comparison. The experimental data in Figure 6-16 are discretized into

two distinct bands. This stems from the necessity to throttle the recirculation valve for

operational reasons. Specifically, as the turbocharger was brought online by partially closing the

exhaust valve (VEXH in Figure 1-1), the recirculation flow rate would climb much more quickly

than the LPT flow rate. Left unchecked, this would result in excessively high recirculation ratios

for the combustor being used, and subsequent engine shut-downs. Thus, restriction of the

recirculation flow was essential to keeping the engine running stably until the turbocharger came

up to speed. Once this was accomplished, the transition from partially closed to fully open was

promptly executed. For data sets in which VREC WAS about 50% open or less, the recirculation

line was considered as "restricted", while for all other data sets, it was considered "unrestricted".

This discretization of data into only two groups for a valve with continuous operation was










appropriate, as this particular valve was very seldom between half-open and full-open. The two

trends exhibit the same shape as each other, but are shifted apart. This is due to the dependence

of R upon the pressure drop through the recirculation line, since the restricted values (higher

pressure drop) are translated downward from the unrestricted values (lower pressure drop). The

experimental trends resemble the modeled trend, though they are shifted downward. This

indicates that the model simulated the performance of a HPRTE having an even less restrictive

recirculation line. The higher recirculation flows resulted in R values that could barely be

achieved in practice with the Build 4 combustor. The model also predicted a turnabout in the

trend at higher pressure ratios, though with the Build 4 configuration, these higher pressure ratios

could not be achieved. The experimental results do, however, exhibit the same upwardly

concave shape as the modeled results, indicating that they would likely change direction if higher

pressure ratios were reached.

An important role of semi-closure is to reduce the flow rate of fresh air necessary to

operate the cycle. As more of the working fluid is composed of recirculated exhaust gases, the

demand for fresh air is decreased as shown in Figure 6-17. For higher values of R the

experimental air flow rates appears to remain somewhat constant. These data correspond with

the data sets recorded when moving through the transition into the pressurized regime. As

mentioned above, when VEXH WAS initially closed, the recirculation flow rate rose more quickly

than the LPC pressure ratio, and in turn air flow rate. Essentially, the data in Figure 6-17 are

very similar to the data in Figure 6-16, but with the ordinate and abscissa exchanged. This is

also seen in the modeled air flow rate trend. The modeled trend is double-valued when shown

over R, pointing to the relationship between LPC flow rate and LPC pressure ratio. Again, the










experimental trends have a steeper slope at lower R values, and may also mimic the modeled

trend if the Build 4 configuration could advance further into the pressurized regime.

In addition to reducing the air flow rate, semi-closure also serves to reduce the exhaust

flow rate of the system as well. Beyond decreasing the size of the turbocharger, a lower exhaust

flow rate will also decrease the heat signature of the HPRTE, which can be paramount in some

applications. Since the air flow rate and exhaust flow rate differ very little, the data in Figure 6-

17 can also be used to approximate the exhaust flow rate over R.

Semi-closure also changes the HPRTE combustion environment. Since the combustor

inlet gases are being diluted with exhaust gases, the stoichiometry within the reaction zone

changes with R at a given total flow rate. Figure 6-18 shows how the equivalence ratio increases

with R alone. The experimental trends appear to agree well with results predicted by the model.

Though the modeled recirculation ratios were much higher, the slopes of all three trends are very

similar. The experimental data in Figure 6-18 separate into the two bands, but show some

intermingling of the restricted and unrestricted cases at higher R values. This points to the

dependence of equivalence ratio on other operating parameters as well. Clearly, in Equation

5.29, the equivalence ratio depends on the air and fuel flow rates. However, with all of these

data sets at full power, this is expected to be influenced more by the air flow rate, which appears

in the denominator and was more varied in its values. The equivalence ratio can therefore be

related to another, primary operational parameter, the LPC pressure ratio, as demonstrated in

Figure 6-19. The experimental data in Figure 6-19 tend to deviate less from their respective

bands. The experimental and modeled trends bear a close likeness to the trends in Figure 6-16,

which reflects the similar, inverse relationship with air-flow-rate that R and equivalence ratio

share.









The more pertinent way in which R affects the combustion environment can be measured

in the composition of the exhaust gas. The impact of R on emissions is here quantified by the

concentrations of NO and CO, the least desirable constituents existing in significant amounts.

The concentrations of these two gases in the exhaust mixture are shown in Figure 6-20. As

expected, and is typically the case, a reduction of the NO concentration is compromised by a

corresponding increase in CO. This results from the opposite dependence of their equilibrium

concentrations on temperature and 02 COncentrations. The formation rate of NO scales linearly

with the product of 02 and N2 COncentrations, and with temperature cubed. As R is increased,

the N2 COncentration remains relatively unchanged while the Oz COncentration drops off, as

affirmed by the data in Figure 6-21. The N2 data in Figure 6-21 indicate a slight rise with R, but

regression lines were added to help illuminate the difference in slopes. It is clear that the

decrease of 02 COncentration over R is faster than the rise in N2 COncentration. While this

dilution of gases detracts from the conditions for NO formation, it also diminishes the rate at

which CO can further oxidize to form CO2. COnsequently, the CO concentrations can be seen to

rise with R. Another influential factor is the temperature at which these reactions are taking

place. The effects of lessening the Oz COncentration are two-fold in that the flame temperature in

the primary zone of the combustor is also reduced. This reduction in temperature further serves

to suppress NO formation, but again, this effect will also reduce the reaction rate to finish

oxidizing CO.

An unexpected phenomenon is evident in Figure 6-20. At the high end of the R range, the

NO concentration takes a marked upward swing. This turnabout cannot be rationalized with any

of the above arguments, since, at still higher R values, all of the NO-diminishing effects are

increasingly present. The most likely reason is related to a minor deficiency in the HPRTE









hardware. The original rover combustor was modified to accommodate recirculation flow by

welding straps over some of the dilution holes. This forces more of the oxygen-dilute gases

through the combustor primary zone for combustion with the intention of restoring oxygen to the

reaction. While this first modification worked well for moderate R values (R < 1.3), it proved to

be insufficient in redirecting 02 at higher R values. As R was increased, the stoichiometry in the

primary zone would progressively become leaner. Instabilities of the flame would then follow,

viewed as violent flickering in the combustor view port, until the flame relocated downstream to

the dilution holes, where more oxygen was present. After the flame relocated, the geometry of

the combustion environment changed, with a greatly increased time for NO forming reactions to

take place.

A particular benefit of cooling the recirculated exhaust gases is the opportunity to extract

the resulting water condensate. However, before this can be achieved to an appreciable extent,

the exhaust recirculation mixture must first be cooled to its dew-point. This was accomplished

with the VARS heat exchangers and the WGC, and relied on the performance of each.

Unfortunately, the VARS employed for these experiments was somewhat undersized to provide

the refrigeration necessary for respectable water extraction.

Another non-dimensional parameter is introduced to present the water extraction data: the

mass flow rate of extracted water, normalized by the fuel mass flow rate. This Water to Fuel

Ratio (WFR) is plotted against the LPC pressure ratio in Figure 6-22. The theoretical maximum

for WFR on a volume basis is around one, which, on a mass basis, equates to approximately 1.2,

though these values can only be approached at high recirculation ratios. Taking this into

consideration, it can be said that the water extraction performance was respectable at lower LPC

pressure ratios. However, the WFR subsequently drops as the tests moved further into the










pressurized regime. The modeled results predicted a much higher WFR throughout the

pressurized regime. However, the model also simulated an ideal VARS equipped with sufficient

refrigeration capacity to maintain a constant, 59 OF (519K) CGC exit temperature. Thus the

modeled WFR depended only on the gas properties at the CGC inlet, and not on VARS

performance.

There are two competing effects relating to HPRTE operation that most directly contribute

to WFR: R and LPC pressure ratio. Greater concentrations of water in the CGC inlet gas

mixture will lead to a greater potential for water extraction by raising the dew-point temperature.

The recirculating gas by itself has a high water concentration, but is dried considerably when

mixed with the fresh air from the LPC. Thus, the greater values of R can be expected to yield

higher WFR values, and R drops with increasing pressure ratio, following the trend seen in

Figure 6-22. However, the effect of increasing the pressure in the recirculation line (increasing

the LPC pressure ratio) should raise the saturation pressure of the water vapor along with its

dew-point, initiating water extraction earlier in the cooling process. Regardless of these two

factors contending with one another, the fact remains that the VARS and heat exchangers must

perform well for appreciable water extraction to be realized. The overriding factor that governed

water extraction performance was undoubtedly the suffering performance of the VARS and

WGC at higher LPC pressure ratios. The heat load on the three recirculation line heat

exchangers is related to the LPC pressure ratio by the recirculation flow rate, which was driven

upward with LPC pressure ratio by throttling VEXH, aS shown by the data in Figure 6-23.

The recirculation gas mixture did reach its dew-point in the CGC most of the time, but by

this point the refrigeration capacity was usually exhausted. Although there were no means

available to directly measure the humidity within, the saturated state of the gas was made evident

























































15 16


by noteworthy quantities of liquid water finding its way through leak-paths in the ducting of the

HPC. Figure 6-24 illustrates how the dew-point temperature changed with the LPC pressure


ratio. Figure 6-24 also shows the increases in CGC inlet and exit temperatures associated with


further pressurization of this HPRTE rig. It is apparent that, thought the dew-point temperature


did increase with LPC pressure ratio, the evaporator temperatures increased at a greater rate.


The water extraction rate should roughly scale with the difference between the inlet dew-point


temperature and CGC exit temperature. Although the CGC temperature difference increased at


higher pressure ratios, the CGC exit temperature still climbed quickly enough to approach the


dew-point temperature, lessening the water extraction rate.


1++
14









09 1 11 1 4

LPC Prssr Rai
+ oe hrmlEfcec

Fiue61 culntpwrotu rmBid4egn us
















































































46 0
Thousands


Figure 6-3. Scaling parameters versus corrected speed.


y =-5E-06x' +0.421x -9176
R2 =0.462






1~-i~y =-3E-07x2 0.038x 950




y =-1E-09x2 + 0.0001x 2.96
R2 =0.292


Em a
m m


445 450 455 46
Thousands


0 21



0 19










011




01


30


n. 5


an


an








4


435 440


N* (rpm)

P*APu Efficiency



Figure 6-2. Corrected Power, power, and thermal efficiency over corrected speed.


1 05

1 00
43 0


43 5 44 0 44 5 45 0 455


N* (rpm)


















y =-0.062x2 + 0.277x + 0.80
R2 =0.812


1 07








1 04*


1 03


1 02



1 0 1 1 1 2 1 3 1 4 1 5 1 6
LPC Pressure Ratio




Figure 6-4. Theta versus LPC pressure ratio.





**



y =-42.2x2 + 129x -23.0
R2 =0.561













10 12 14 16 18 20
LPC Pressure Ratio

*P*Modeled



Figure 6-5. Power versus LPC pressure ratio.















*430



o=-.7X2 + 3.27x -1.12
. R2 =0.947 380



y =18.9x2 -100x +442 33
R2 =0.802



as




230


20








1 6





0.2


CO 0


10 11 12 13 14 15
LPC Pressure Ratio

Mass FLow Rate a HPT nT A HPC AT



Figure 6-6. High pressure turbomachinery temperature differences and mass flow rate versus
LPC pressure ratio.







Slope =84.33 p












Slope = -6.44





10 11 12 13 14 15 16
LPC Pressure Ratio

aMass Flow Rate + HPT nT A HPC AT



Figure 6-7. Percent change in high pressure turbomachinery temperature differences and mass
flow rate versus LPC pressure ratio.

















17



15




12

E11













35 -


46 0
Thousands


15 -- 7

2


25 3 35 4


Figure 6-9. Generalized effects of HPC pressure ratio and CTR on specific power, illustrated on
a T-S diagram.






















E 47

S46


S45


0 44

0- 43


2.60-2.66
2.69-2.72
2.73-2.76
2.81-2.83
2.85-2.87
2.89-2.91


403





Figure 6-10.


335


345


355


CTR


Specific power versus CTR, for various HPC pressure ratios.


o


2.60-2.66
2.69-2.72
2.73-2.76
2.81-2.83
2.85-2.87
2.89-2.91


S56


0- 54


335


345


355


CTR


Figure 6-11. Corrected power versus CTR for various HPC pressure ratios.














































































355


y =1E8x-08x .0009x + 21.4
R2 =0.492

y =-2E-09x2 + 0.0004x 8.84
R2 =0.958


30


27



25
4


435


440


445


450


460
Thousands


IM* (rpm)


+ HPC Pressure Ratio a CTR



Figure 6-12. HPC pressure ratio and CTR versus corrected speed.




O 150



0 145
y =-0.327x2 + 2.30x -3.91
R2 =0.658





0+ *






0 125



0 120
325 330 335 340 345 350

CTR




Figure 6-13. Thermal efficiency versus CTR.












































































5 46 0
Thousands


0 150
y =-0.084x2 + 0.505x 0.617
R2 =0.349
0 145 '





0 135





0 130



*



0 120
2 55 2 60 2 65 2 70 2 75 2 80 2 85
HPC Pressure Ratio




Figure 6-14. Thermal efficiency versus HPC pressure ratio.





72 5
y = -1E-07x2+ 0.011x -201
R2 =0.576
S72 0









4304 404 50


2 90 2 95


N* (rpm)



Figure 6-15. HPC isentropic efficiency versus corrected speed.


45



















y =5.38x2 1.8x + 14.4
R2 =0.913










13
y =0.808x2 -2.91x + 3.74
R2 =0.882



10 12 14 16 18 20
LPC Pressure Ratio

Restricted a Unrestricted Modeled



Figure 6-16. Recirculation ratio versus LPC pressure ratio.





1 00


0 00
I~y = 0.795x2 -2.71x + 2.95
P R2 =0.975







0 40
11 1 3 1 5 1 7 1 9 2
rrR
+ etice Ursritd Moee

Fiue61.Feharfo at essrcruainrto













0 60


y =0.032x2 + 0.213x + 0.089

0 45 ~R2 =0.971



y =0.358xo.1
0 40 + *R2 =0.643





11 12 13 14 15 16 17 18 19 20


Restricted a Unrestricted Modeled



Figure 6-18. Equivalence ratio versus recirculation ratio.


O 60















m "y = .649-1.178
~R2 =0.617



y =0.337x2 1.13x + 1.32
R2 =0.959


10 12 14 16 18 20
LPC Pressure Ratio

+ Restricted a Unrestricted Modeled



Figure 6-19. Equivalence ratio versus LPC pressure ratio.













45


40


35

E
P. 30



20



O 1


10


07 09 11 13 15 17 19


NO mCO



Figure 6-20. Nitric oxide and carbon monoxide concentrations versus recirculation ratio.


ge 12



o



On 1 0



9


07 09 11


13 15 17 19
R


Figure 6-21. Oxygen and nitrogen concentrations versus recirculation ratio.

























































+ *+ *






+ ~y = -1.98x2 +5.57x 2.91
+ R = 0.840










11 12 13 14 15
LPC Pressure Ratio


04




y =3.27x2 -9.14x +6.50
t \=R2 =0.691









10 12 14 16 18 20
LPC Pressure Ratio

+ WFR *Modeled



Figure 6-22. Water-to-fuel ratio versus LPC pressure ratio.


12



11







do

u, 08


Figure 6-23. Recirculation flow rate versus LPC pressure ratio.
















y =74.6x2 -64.8x + 113
17n R2 =0.811



S130 m cL~

gy =3 -12.5x2 + 59.6x + 47.3
a* R2 =0.968





Sy =-56.7x2 + 217x -95.5
5o R2 =0.77



10 11 12 13 14 15 16
LPC Pressure Ratio

CGC Exit a CGC Inlet r Dew-Point, CGC Inlet



Figure 6-24. Temperatures related to water extraction versus LPC pressure ratio.









CHAPTER 7
CONCLUSIONS AND RECO1V1VENDATIONS

Summary

Following these experimental efforts, several conclusions can be drawn about the overall

success of the study, and about the theoretical capabilities offered by the HPRTE/VARS

combined cycle. Throughout this research, four engine configurations were tested and

progressively altered in order to arrive at an experimental apparatus best suited for operating the

HPRTE/VARS, and for demonstrating its qualities. It was shown that the VARS hardware can

be successfully integrated with pressurized and semi-closed engine ducting, and driven by the

hot recirculating gases, resulting in a synergistic combination of cycles. The successful

operation of the HPRTE was also demonstrated, as the HPRTE was pressurized and semi-closed

to various degrees. Despite operational limitations presented by certain components, data were

acquired and adjusted to successfully prove the following theoretically proposed concepts:

* The HPRTE power output can be increased/decrease by diverting exhaust flow
through/around the LPT, varying the pressure ratio across the turbocharger, and the total
mass flow rate through the engine.

* The HPRTE efficiency can remain approximately constant for a variety of power levels,
since the dimensionless operating point of the engine changes very little, or not at all
during power transitions.

* The emission of NOx from the HPRTE is reduced as R is increased.

* The mass flow rates of air into and exhaust gas out of the HPRTE are reduced as R is
increased.

* Significant quantities of water can be condensed and extracted from the recirculating
gases.

The performance of the HPRTE/VARS can be hindered by certain of its components. The

pressure drop in the recirculation line was the primary detraction from the power-making

potential of the engine. On this engine rig, the pressure difference between the HPT exit and









HPC inlet significantly reduced the HPT pressure ratio, raising the HPT exit temperature near to

its limit. When this occurred, there was little or no margin remaining in HPT temperatures to

apply a load using the dynamometer. Gas leaks in the ducting and engine casing were also

suspected to have an adverse effect on net power, though the extent to which they detracted from

the total flow rate remains uncertain.

While an ideal engine would operate at a constant speed throughout the pressurized

regime, this engine, equipped with a mechanical governor, was subj ect to a reduction in speed as

the engine was pressurized. The speed-droop was small, but discernable. A small decrease in

corrected flow rate was observed at the engine speed lessened, as well as a drop in both the HPC

and HPT pressure ratios. With the drop in pressure ratios followed a drop in both temperature

ratios, and in turn the temperature differences (proportional to net power) across the HPC and

HPT. Despite the departure from its design speed, the HPRTE/VARS suffered only a small

decrease in thermal efficiency, barley evident in the data scatter.

The VARS described in Chapter 3 was designed specifically for this application, but was

intended for a slightly different ducting configuration, feeding cooler gases to the HGC. In

actuality, the HGC inlet gases were approximately 300-400 OF (167-222 K) hotter than designed

for, increasing the HGC effectiveness. This effectively shifted the optimal VARS operating

point to much lower HGC and CGC flow rates, and caused the VARS to be overwhelmed at

higher heat loads. This effect prohibited significant water extraction rates from being realized at

higher LPC pressure ratios, because the CGC temperatures rose much faster than the dew-point

temperature. Substantial water extraction was however accomplished at lower LPC pressure

ratios where R was highest and the VARS performed respectably.









Recommendations

Through the course of these HPRTE/VARS experiments, much was learned about the

positive attributes of the combined cycle, and about its limitations. The range of operations and

performance of this combined cycle were found to be limited by several details associated with

its hardware. Difficulties and uncertainties were also encountered while completing the data

reduction and interpretation due to inadequacies in the instrumentation. These experiences have

served to generate several recommendations that would improve the quality of future

HPRTE/VARS experimental endeavors.

Significant improvements can result from utilizing a VARS having a wider range of

operations. This is believed as key since the HPRTE itself is capable of operating with a variety

of flow rates, temperatures, and pressures in the recirculation flow path. An ideal VARS would

posses more capacity and a more sophisticated control scheme to maintain a constant HPC inlet

temperature, and perhaps excess refrigeration capacity. This would most strongly impact the

water extraction rate from the HPRTE. While the HPC inlet (CGC exit) temperature remains

constant, the dew-point temperature would increase with LPC pressure ratio, increasing the water

extraction rate. The production of excess refrigeration from the VARS would also be a favorable

quality, and would allow for a new dimension of operations to be investigated. Specifically, this

would permit experiments where excess refrigeration can be traded for engine efficiency and

water extraction, or vice versa.

The potential of an HPRTE to make power could be greatly increased if the pressure drops

in the recirculation line are kept minimal. The VARS heat exchangers used in these experiments

were relatively compact, and it would be tolerable in research-oriented applications if they were

somewhat larger. This would serve to reduce the back-pressure on the HPT and restore some

margin in the HPT exit temperature for engine loading. Another point worth noting is that










engines with higher pressure ratios should suffer less from a given differential pressure across

the recirculation line. These experiments saw engine recirculation differential pressures of up to

4 psi (28 kPa), which resulted in a 20% reduction in the HPT pressure ratio. The same

differential pressure with a higher HPT pressure ratio should be a less significant detraction from

the HPT temperature difference.

The range of HPRTE operation could also be markedly enhanced by employing a variable

geometry turbocharger (VGT). VGTs have the ability to vary the AR ratio on their turbine

housings, allowing for a variety of LPT flow rates for a given LPT pressure ratio. This would all

but eliminate the challenge of selecting one turbocharger with a constant AR ratio to match the

variety of recirculation ratios and LPC pressure ratios seen by the HPRTE. A VGT would also

most likely eliminate the need for a recirculation valve. The recirculation valve has functioned

primarily to restrict recirculation flow after the wastegate (VEXH) is initially closed, diverting

more flow through the LPT. When operating in this way, the recirculation pressure drop is

higher than necessary. A VGT would function with a larger AR ratio during initial

pressurization, and may further do away with the need for a wastegate. It may be preferable,

however, to retain the service of a wastegate as it would allow multiple R-versus-pressure-ratio

lines on which to operate.

Future HPRTE configurations may also benefit from additional water extraction points

directly upstream of the HPC inlet. Condensate was observed on several occasions leaking from

flanges in this location, confirming the presence of liquids in the ducting near the HPC inlet.

The gas entering the HPC can be saturated with water vapor (ideally it always is) and has even

been observed as "misty" with tiny, suspended water droplets. While the small droplets entering

the HPC can be a performance enhancement, they can also present a problem if they coalesce on










the ducting walls, or become too large for the HPC rotor being used. This can result unnecessary

erosion of the compressor wheel.

The last recommendation following this work relates to instrumentation, and would be

found most useful in applications where the water extraction potential is exploited-the inclusion

of a duct-mounted humidity sensor. This instrument would significantly increase the ease of

data reduction, eliminating many of the approximations and assumptions discussed in Chapter 5.

A more reliable knowledge of this quantity would result in a much more accurate dew-point

temperature, heat exchanger effectiveness, and overall quantification of water extraction

performance.









APPENDIX A
OPERATING PROCEDURES FOR HPRTE/VARS


Starting Procedure:

Prior to Start:

1 Tumn on the VCR Recorder, tumn VHS Camera to CAMERA .

2. Do a communications check.

3. Plug in ARU. Place the ARU in standby by rotating spring return switch clockwise. The
YELLOW light should come on.

4. Tumn on the condensate removal pumps.

Engine Start:

1. Tumn on the DC circuit breakers local to the skid. Red indicator light should come on
locally and in the Control Room

2. Move the Speed trim Valve to half way between open and shut.

3. Check open the High Pressure Fuel Manual Isolation Valve. Valve handle cross-line with
the fuel pressure line.

4. Move S1 to ON.

5. Check Oil Pump #1 to Bypass.

6. Check Fuel Pump to Bypass (S9).

7. Check Oil Pump #2 to Normal.

8. Tumn Oil Pump #2 to On. Confirm pressure light on before continuing.

9. Check fuel pump Off.

10. Check Air Pressure light on.

11. Move Ignition switch to ON. This enables the ignition system and the starter

12. Check the 12VDC starter Off.

13. Check Spray Cooler switch Off.










14. Check the Combustor Air Valve Off.

15. Check the solenoid Fuel Valve Off.

16. Press the RUN button. The lighted button and the control room indicator light should
come on.

17. Press and Hold the START button. Check Fuel Pressure light on. The fuel valve solenoid
energized light should be Off.

18. At 300 PRM (or 40 PSIG fuel pressure), turn the Solenoid Fuel Valve on. Verify ignition
and increase in engine speed.

19. Release the start button when green light (starter solenoid energized) goes out.

20. Engine speed should increase to self-sustaining, about 2700 dyno RPM. Wait until fuel
has burned off in the sight glass, if necessary. Slowly move the Speed Trim valve to full
OPEN, rotated down. Engine speed should increase to approximately 3000 RPM as read
on the Dyno.

21. Tumn the DC exciter and Local switch to Off. Verify local oil pressure.

Recirculation, ARU, Dyno, and Turbocharged Operation:

1. Open ARU V14. ARU V13 can be opened 1 V/2 tumns to hasten warm up.

2. Slowly open the Recirculation Isolation Valve to until recirc DP=1.6.

3. Partially close the Main Air Inlet Valve. MAI DP=0.5 in.-H20 oil max. Recirc DP=6.5".
Recirc ratio=0.6.

4. Close the Dyno outlet valve until the flow is about 5 GPM and pressure is 28 to 30 PSIG.

5. ARU System pressure will rapidly build up. When the ammonia receiver pressure reaches
210 psig, open the condenser cooling water valve to maintain pressure around 235 psig,
but below 250 psig. Run data shows about 4 gpm is appropriate at full dyno load (no
boost). Discharge weak solution if required.
6. Adjust solution flow to 3.2 gpm by throttling V13 if required.

7. Slowly adjust the Boost control pressure to 8.0 PSIG. Verify Boost Control Valve stem
position is 28% open. To avoid burner instability use the Recirculation Control Valve to
keep R less than 1.0. Flame transition should occur near this setting.











Recirculation Setting and Steady-State Operations:

1 Adjust HRVG Heat Load to 23TR using Recirc and Boost control valves. During this
time ARU should be adjusted to limit T-HPCI below 90F. Record data as necessary.
Purge non-condensables and remove solution as necessary.

PR LPT may be 1.0.

2. Increase HRVG heat load within stable limits. Record data Column 3 and 4.
Recirculation control and Boost control maybe used to achieve PR LPT up to 1.8.


Shut Down:

1. Remove Dyno load slowly from engine.

2. Open Boost Control valve to 7.5 PSIG. Restrict Recirc control to DP=16" (about 45
degrees). Open MAI.

3. Reduce engine speed to 2900 RPM using the Speed Trim Valve.

4. Close the Manual Fuel Shut Off Valve.

5. Press the STOP button.

6. Move the Fuel Valve switch to Off.

7. Turn Off VCR.


Post Shut Down:

1. Limit soak-back to under 3800F. Motor the engine for 30 seconds as necessary.

2. Turn ARU off. Close ARU valves V14 and V2.

3. Collect Data sheets.

4. Review Tape immediately. Synchronize data sheets and data stream with unusual events.

5. Critique and review the test run. Check for signs of fluid leaks. Reconfigure the engine
for a quick restart.










APPENDIX B
SETUP PROCEDURES FOR HPRTE/VARS


Support Systems:

Site Chill Water:

1. Slowly open the source and return overhead isolation valves.

Cooler Chill Water:

1. Connect the process water hoses to the cooler.

2. Supply water to the coolers by opening all isolation valves.

3. Verify main cooler flow by listening for flow noise. Record initial flows:

4. Record minor leaks for later resolution.

ARU Chill Water:

1. Connect the process water hoses to the ARU.

2. Supply water to the cooler by opening all isolation valves.

3. Verify main cooler flow by listening for flow noise.

4. Record minor leaks for later resolution

Boost Control Valve and Waste Gate Control Air:

1. Set supply air regulator by the South door to 40 psig.

2. Ensure the Supply, Boost Control and Waste Gate Bleed toggles are closed.

3. Ensure the Waste Gate Isolation toggles are open.

4. Verify the Supply Air Pressure at the control panel is between 25 and 30 psig by
adjusting the South door regulator in 1. above.

5. Open the Fisher Control Regulator until the Boost Control Air gage reads 15 psig. This
opens the 6" Boost Control valve completely.

6. Visually verify the position of the Boost Control Valve is full open by observation of the
valve stem indicator.









7. Verify the turbocharger Waste Gate Regulator is set to 20 psig.

8. Verify operation of the Turbocharger Waste Gate with an assistant listening for actuation,
from full open to full closed, by closing the Waste Gate Isolation Toggle and opening the
Waste Gate Bleed Toggle.

9. Reset the toggles for Waste Gate position Full open : Isolation Toggle Open/Bleed
Toggle closed.

12 Volt Battery Check

1 Unplug the battery charger and store the charger.

2. Throw the isolation switch.

3. Verify each start battery has a cold reading of 13.2 volts minimum. This ensures that the
batteries are fully charged. Record Voltage:

4. Verify each ignition battery has a cold reading of 12.6 volts minimum. This ensures that
the batteries are sufficiently charged. Record Voltage:



Fuel Supply

1. Verify Gravity and speed trim lines to engine skid. Open the ROVER side of the gravity
system. Place the speed trim valve in the open position, rotated down. Shutting this valve
will trim engine speed and eventually shut the engine down.

2. Purge the fuel Pump accumulator to charge with air and rig for remote valve operation.

3. Using the battery charger AC supply, plug in the fuel transfer pump and verify High and
Low speed operation.

4. Check fuel level by sight glass and control panel gage, minimum %/ full. Fill as necessary.

5. Fill the seven gallon tank under the fuel cabinet, if required.

6. Check the entire fuel cabinet and hoses for leaks.

7. Have a full 5 gallon fuel can standing by in the fuel closet, if required.

8. Dry and position the skid drip pan and the fuel drain drip pan.

Oil Levels









1. Verify Rover engine oil level. Refill with single viscosity 10W oil as necessary up to half
way between the high and low markings.

2. Safety wire the dipstick.

3. Verify Turbocharger oil level. Fill with 10W-30 oil as necessary as not to flood the turbo
oil scavenge port. Visually check the condition of the oil.

Dyno Setup and Oil Cooler Process Water

1. Crack open the Turbocharger Lube Oil Cooler water supply tap.

2. Route system drain lines outside under the main overhead door.

3. Fully open the water supply to the Froude Dyno. Purge Dyno.

4. Verify the in-line Rota meter is reading 6.9 GPM. Adjust Dyno inlet and outlet valve to
full open.

5. Verify the water brake is fully unloaded, that is, the geared handle is full to the CCW
position.

6. Verify the gear lock remains disengaged.

7. Visually confirm gland leakage.

8. Confirm discharge flow from four drain lines. They are; 1.)The Rover oil cooler, 2.) The
dyno, 3.) The dyno drip pan and 4.) The Turbocharger Lube Oil Cooler.

Water Recovery

1. Attach the Water Recovery Reservoir to the load cell.

2. Supply AC power to the drain pumps and turn them on. Verify rotation










Inlet, Recirculation and Exhaust Start-up Check

1. Check the exhaust system to ensure all penetrations are covered and j points are tight.

2. Verify that the Boost Control Valve is fully open.

3. Verify that the Rover Inlet Isolation Valve is fully open

4. Verify the Waste Gate Valve is fully open.

5. Verify the Rover Recirculation Valve is fully shut.

Engine room Preparation

Room Ventilation:

1. Open the main bay door about five feet (to marked line).

2. Turn on the lab ventilation fan and the compressor room fan.

3. Turn the air conditioning thermostat bypass, at the north wall, to "on".

Lab Over-watch:

1. Check the lab area for debris that could be ingested into the engine or present a tripping
hazard.

2. Attach the Safety Chain at the hall outside the Lab.

3. Move fire extinguishers to areas in the lab where they are readily accessible.

4. Visitor Policy: all visitors should be checked in, briefed, and supplied with safety
equipment before the run set-up begins. Optimally, all visitors should be supplied with
Listen-Only communication head gear. No late or unannounced visitors are allowed.

5. Put the scatter shields in place for turbocharged runs, if required.

Gas Analysis and Data Acquisition

1. Refer to the gas analysis set up procedure, separate from this document.

2. Verify that all thermocouples are reading properly by both the analog and digital data
acquisition systems. Confirm conformance to the instrumentation map. This should be
completed a day in advance of the run.









3. Verify that all pressure taps are reading properly by both the analog and digital data
acquisition systems. Confirm conformance to the instrumentation map. This should be
completed a day in advance of the run.

4. Verify the analog pressure, temperature, and manometer reading legends are clearly
displayed on the panel near the instruments.

5. Record the initial reading from the Water Recovery load cell on the Data Sheet.

ARU Setup


1. Check if the solution receiver is at least one quarter full. If it is lower than 1/4, see
troubleshooting guide.
2. Check all the valve positions according to valve tag list.
3. Check closed V13 wide (pump bypass).
4. Record the initial Solution Receiver tank level.

5. Record the initial NH3 Receiver Tank level.

6. Set V14 shut (pump discharge).
7. V15 at 3-tumn open, Set by Supplier (Mattingly) (col feed).
8. Adjust water valves V26 and V27 to get 2 gpm in condenser and 10 gpm in absorber.
Record initial flows:
9. Tumn power switch on. Panel lights should go from Red to Yellow until HRVG inlet air
becomes hot. Then the Green light will come on starting the solution pump and opening
the solenoid valves.
Video and Audio Recording

1. Set up the VCR in the lab. Run the cables to the control room to the video monitor.

2. Set up the microphone to record the communications loop. Hook this into the VCR sound
mnput.

3. Insert a new VCR tape for the day's activity. Ensure tape is recording on E.P. (extended
play).

4. Synchronize TIMVE and DATE.

5. Complete a system check to be sure that all the monitoring and recording systems are
working correctly. This system is used to verify the data set switch points and aids in
improving each subsequent runs through lessons learned.

Personnel Safety Equipment and Communications











1. All personnel should wear appropriate clothing for an environment where high
temperature piping, heavy equipment and high speed rotating equipment exist, i.e., long
sleeve shirt, long pants, closed-toe and -heel shoes, and no loose fitting items or jewelry.
TURN OFF CELL PHONES.

2. Check all the communication gear. Check batteries. All units should be on the same
channel and in Push-To-Talk (PTT) mode. All units should be on TX, not INTT. Use
channel A.

3. All personnel and visitors should have hearing protection, either communication sets or
ear muffs.

4. All personnel and visitors should have eye protection.































































1400
48.0
1.32


1419
50.3
1.32


APPENDIX C
ADJUSTED EXPERIMENTAL DATA

Provided below is the reduced data for all of the successful Build 4 engine runs. Each

engine run is identified as B4-X, where X represents the run number. Each run had between one

and five separate trials. The reduced parameters shown were calculated after the data

adjustments were made, according to the discussion in Chapter 5.


Table C-1. Data from runs B4-1 and B4-2.
Run B4-1


1


B4-2
2 1


Tnial
Main Air Inlet
Ambient Temperature (F)
Ambient Pressure (psia)
Ambient Humidity Ratio
Mass Flow Rate (lbm/s)
Low Pressure Compressor
Inlet Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Efficiency (%)
Pressure Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
High Pressure Compressor
Inlet Temperature (F)
Inlet Pressure (psia)
Adiabatic Exit Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Efficiency (%)
Pressure Ratio
Total Gas Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Combustor
Fuel Flow Rate (lbm/s)
Phi
Combustion Efficiency
Combustor Exit Temperature (F)
High Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma


2 3


69.3
14.7
0.0189
0.00

96.0
1.40
2.8
54.9
1.06
0.52
0.54

76.9
15.2
330.4
1.38
115.4
70.8
2.85
1.31
1.29

0.018
0.50
91.7
1472


70.7
14.7
0.0189
0.00

96.0
1.40
4.3
54.8
1.08
0.54
0.56

77.7
15.6
330.3
1.38
119.2
71.3
2.86
1.35
1.30

0.019
0.49
91.7
1463


84.9
14.7
0.0171
0.00

95.6
1.40
8.9
66.6
1.18
0.64
0.66

76.6
17.1
327.3
1.38
131.0
71.5
2.82
1.49
1.31

0.020
0.44
96.6
1485

1414
48.2
1.32


86.5
14.7
0.0174
0.00

97.3
1.40
7.9
71.0
1.18
0.60
0.62

79.7
17.1
329.9
1.38
130.9
71.7
2.81
1.49
1.31

0.020
0.47
96.6
1474


87.8
14.7
0.0174
0.00

98.7
1.40
13.0
73.3
1.28
0.68
0.70

89.2
18.5
339.2
1.38
139.4
70.2
2.72
1.59
1.30

0.020
0.43
97.3
1480


1429
43.3
1.31


1430
44.4
1.31










Run
Tnial
Shaft Power (hp)
Isentropic Efficiency (%)
Expansion Ratio
Corrected Flow Rate (lbm/s)
Recuperator
Hot Side dT (F)
Hot Side Heat Rate (Btu/s)
Cold Side dT (F)
Cold Side Heat Rate (Btu/s)
Heat to Ambient (Btu/s)
Effectiveness
Low Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Efficiency (%)
Expansion Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Hot Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Warm Gas Cooler
Gas Side dT (F)
Water Side dT (F)
Gas Side Heat Rate (Btu/s)
Water Side Heat Rate (Btu/s)
Heat Loss to Ambient (Btu/s)
Effectiveness
Cold Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Overall Performance

Power (hp)
Thermal Efficiency (%)
Average Water Extraction Rate (gph)
Average Fuel Flow Rate (gph)


B4-1
1
-180.0
85.0
2.85
0.82

-176.6
-78.5
218.3
72.6
-5.9
0.32

899
17.5
1.33
-2.8
35.8
1.19
0.36
0.49


B4-2
1
-204.1
85.0
2.82
0.84

-120.4
-88.5
217.2
82.5
-5.9
0.32

938
19.4
1.34
-8.9
53.2
1.32
0.47
0.58


2
-186.4
85.0
2.86
0.83

-165.9
-82.6
223.2
76.7
-5.9
0.32

910
17.9
1.33
-4.3
45.5
1.21
0.39
0.52


2
-202.1
85.0
2.81
0.85

-122.5
-85.7
210.0
79.9
-5.8
0.31

926
19.1
1.34
-7.9
52.1
1.30
0.45
0.57


3
-211.4
85.0
2.72
0.86

-136.1
-89.9
207.4
84.2
-5.7
0.32

937
20.5
1.34
-13.0
58.7
1.39
0.52
0.61


-427.7 -431.6 -431.8 -412.6 -410.4
-91.4 -94.1 -98.8 -98.0 -99.9


-393.4
11.0
-86.8
59.1
-27.7
0.73


-389.9
12.1
-85.6
61.5
-24.1
0.71


-405.6
12.5
-91.9
65.8
-26.1
0.70


-398.5
13.3
-93.6
69.1
-24.4
0.69


-405.3
14.1
-96.4
73.2
-23.2
0.68


-49.9 -52.0 -58.6 -61.2 -58.3
-23.1 -22.2 -25.1 -25.9 -25.2


1.53
64.5
13.90
5.81
9.21


1.51
67.2
13.97
4.00
9.54


1.35
73.1
14.40
2.88
10.08


1.48
71.2
14.05
2.69
10.06


1.35
72.0
13.64
1.86
10.47




























































-127.4
-75.9
209.4
69.9
-6.0
0.30


-140.3
-91.6
196.4
85.7
-5.9
0.32


-143.8
-93.7
202.1
87.9
-5.8
0.32


1423 1437 1406 1398 1416
52.3 52.2 43.1 52.4 52.3
1.32 1.32 1.32 1.32 1.32
-220.4 -218.1 -181.7 -219.7 -216.8
85.0 85.0 85.0 85.0 85.0
2.70 2.65 2.87 2.65 2.60
0.86 0.86 0.84 0.87 0.87


Table C-2. Data from runs B4-3 and B4-5.
Run B4-3


1


B4-5
2 3 1


Tnial
Main Air Inlet
Ambient Temperature (F)
Ambient Pressure (psia)
Ambient Humidity Ratio
Mass Flow Rate (lbm/s)
Low Pressure Compressor
Inlet Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
High Pressure Compressor
Inlet Temperature (F)
Inlet Pressure (psia)
Adiabatic Exit Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Total Gas Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Combustor
Fuel Flow Rate (lbm/s)
Phi
Combustion Effciency
Combustor Exit Temperature (F)
High Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Expansion Ratio
Corrected Flow Rate (lbm/s)
Recuperator
Hot Side dT (F)
Hot Side Heat Rate (Btu/s)
Cold Side dT (F)
Cold Side Heat Rate (Btu/s)
Heat to Ambient (Btu/s)
Effectiveness


2

87.0
14.7
0.0169
0.00

97.8
1.40
17.1
78.8
1.39
0.70
0.73

108.3
20.1
355.9
1.38
147.9
69.4
2.60
1.70
1.30

0.021
0.43
97.3
1458


85.2
14.8
0.0186
0.00

95.0
1.39
16.7
72.4
1.34
0.73
0.75

93.6
19.3
342.4
1.38
145.0
70.7
2.70
1.66
1.30

0.021
0.41
97.6
1477


86.5
14.8
0.0186
0.00

95.0
1.39
18.3
71.4
1.36
0.74
0.77

102.7
19.7
350.6
1.38
144.4
70.6
2.65
1.66
1.29

0.021
0.40
98.0
1495


87.2
14.8
0.0186
0.00

95.0
1.40
2.7
39.4
1.04
0.50
0.51

71.5
15.0
326.3
1.38
116.9
70.9
2.87
1.31
1.30

0.018
0.52
96.1
1514


86.0
14.7
0.0169
0.00

95.5
1.40
15.8
77.7
1.37
0.69
0.71

97.2
19.8
345.9
1.38
149.1
69.5
2.65
1.71
1.31

0.021
0.44
96.2
1435


-128.5
-91.5
201.5
85.5
-6.0
0.32


-132.2
-95.2
210.2
89.2
-6.0
0.32










Run
Tnial
Low Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Efficiency (%)
Expansion Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Hot Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Warm Gas Cooler
Gas Side dT (F)
Water Side dT (F)
Gas Side Heat Rate (Btu/s)
Water Side Heat Rate (Btu/s)
Heat Loss to Ambient (Btu/s)
Effectiveness
Cold Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Overall Performance

Power (hp)
Thermal Efficiency (%)
Average Water Extraction Rate (gph)
Average Fuel Flow Rate (gph)


B4-3
1

950
21.5
1.34
-16.7
61.4
1.45
0.56
0.63

-418.5
-104.0

-412.2
14.8
-100.1
74.0
-26.0
0.69


B4-5
2 3 1


2

941
21.5
1.34
-17.1
61.7
1.46
0.57
0.63

-389.9
-103.5

-435.6
16.3
-112.1
81.8
-30.4
0.70


964
21.8
1.34
-18.3
63.1
1.47
0.57
0.64

-420.4
-102.6

-418.0
15.0
-99.2
75.1
-24.1
0.68


921
17.0
1.34
-2.7
21.6
1.16
0.32
0.46

-409.8
-90.5

-399.8
11.7
-91.8
65.2
-26.6
0.72


923
21.2
1.34
-15.8
61.5
1.44
0.55
0.63

-383.6
-104.4

-424.7
16.2
-112.4
79.1
-33.3
0.70


-51.0 -46.2 -61.4 -46.1 -38.7
-23.4 -21.0 -27.6 -21.5 -18.1


1.28
75.4
13.85
1.97
10.80


1.23
73.7
13.59
1.64
10.76


1.66
64.8
13.81
6.27
9.31


1.49
70.5
12.96
1.67
10.80


1.42
68.9
12.76
1.49
10.71




























































-108.3
-97.1
210.3
91.0
-6.1
0.30


-105.3
-98.0
210.7
92.0
-6.0
0.29


-109.4
-110.9
221.0
104.6
-6.3
0.33


1387 1373 1414 1415 1418
53.6 51.2 52.2 52.5 56.6
1.32 1.32 1.32 1.32 1.32
-226.8 -215.2 -223.4 -224.2 -240.9
85.0 85.0 85.0 85.0 85.0
2.73 2.71 2.71 2.69 2.65
0.86 0.86 0.89 0.90 0.88


Table C-3. Data from runs B4-7 and B4-8
Run B4-7


1


B4-8
2 3 4 1


Tnial
Main Air Inlet
Ambient Temperature (F)
Ambient Pressure (psia)
Ambient Humidity Ratio
Mass Flow Rate (lbm/s)
Low Pressure Compressor
Inlet Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
High Pressure Compressor
Inlet Temperature (F)
Inlet Pressure (psia)
Adiabatic Exit Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Total Gas Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Combustor
Fuel Flow Rate (lbm/s)
Phi
Combustion Effciency
Combustor Exit Temperature (F)
High Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Expansion Ratio
Corrected Flow Rate (lbm/s)
Recuperator
Hot Side dT (F)
Hot Side Heat Rate (Btu/s)
Cold Side dT (F)
Cold Side Heat Rate (Btu/s)
Heat to Ambient (Btu/s)
Effectiveness


85.1
14.8
0.0186
0.00

90.3
1.40
18.1
69.8
1.36
0.73
0.75

88.7
19.7
337.5
1.38
151.2
70.7
2.73
1.73
1.33

0.021
0.42
96.1
1427


86.0
14.8
0.0186
0.00

92.5
1.40
13.3
72.5
1.30
0.65
0.66

92.2
18.9
342.3
1.38
146.0
70.3
2.71
1.66
1.33

0.020
0.46
94.9
1417


87.7
14.8
0.0190


88.1
14.8
0.0200


84.2
14.7
0.0189


0.00 0.00 0.00


93.3
1.40
14.9
73.3
1.33
0.67
0.69

92.3
19.3
339.4
1.38
147.1
71.0
2.71
1.69
1.33

0.023
0.49
96.8
1521


94.1
1.39
15.6
73.3
1.34
0.68
0.70

93.3
19.5
340.1
1.38
148.1
70.8
2.69
1.70
1.33

0.024
0.50
97.2
1545


90.1
1.39
27.7
69.6
1.48
0.86
0.88

89.5
21.4
331.7
1.39
157.5
70.5
2.65
1.85
1.31

0.023
0.38
99.6
1473


-98.3
-96.1
203.3
89.8
-6.3
0.33


-98.5
-90.6
198.9
84.4
-6.2
0.33










Run
Tnial
Low Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Efficiency (%)
Expansion Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Hot Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Warm Gas Cooler
Gas Side dT (F)
Water Side dT (F)
Gas Side Heat Rate (Btu/s)
Water Side Heat Rate (Btu/s)
Heat Loss to Ambient (Btu/s)
Effectiveness
Cold Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Overall Performance

Power (hp)
Thermal Efficiency (%)
Average Water Extraction Rate (gph)
Average Fuel Flow Rate (gph)


B4-7
1

947
21.6
1.34
-18.1
64.4
1.46
0.57
0.64

-391.9
-104.1

-412.5
16.1
-105.1
78.6
-26.4
0.66


B4-8
2 3 4 1


937
20.3
1.34
-13.3
62.7
1.38
0.51
0.60

-390.7
-105.3

-411.0
15.8
-106.1
78.6
-27.5
0.67


964
20.8
1.34
-14.9
63.4
1.41
0.52
0.61

-394.5
-107.5

-437.4
15.8
-114.1
87.2
-26.9
0.69


970
21.0
1.34
-15.6
63.9
1.42
0.53
0.62

-393.2
-107.6

-441.0
16.2
-115.5
92.3
-23.2
0.69


971
23.9
1.34
-27.7
67.6
1.62
0.65
0.67

-413.8
-109.5

-422.0
16.3
-107.1
79.2
-27.9
0.67


-71.0 -58.9 -59.5 -60.6 -76.8
-32.0 -25.5 -26.8 -27.5 -37.7


1.37
75.6
13.72
0.81
10.94


1.57
69.1
13.04
0.66
10.52


1.52
76.3
12.77
0.96
11.86


1.51
76.1
12.39
0.94
12.20


1.16
83.4
14.13
1.24
11.71



























































-95.4
-94.2
199.9
88.2
-6.1
0.31


-99.4
-97.3
210.8
91.1
-6.2
0.30


-99.0
-98.3
210.9
92.2
-6.2
0.30


1354 1363 1409 1403
52.5 52.4 54.4 55.0
1.32 1.32 1.32 1.32
-225.0 -223.4 -225.9 -227.5
85.0 85.0 85.0 85.0
2.75 2.73 2.75 2.75
0.88 0.88 0.85 0.85


Table C-4. Data from runs B4-10 and B4-11.
Run B4-10


1


B4-11
1

83.8
14.8
0.0169
0.00

91.5
1.40
16.8
67.7
1.34
0.69
0.71

88.8
18.9
338.7
1.38
148.5
71.2
2.75
1.69
1.29

0.022
0.47
97.4
1500


Tnial
Main Air Inlet
Ambient Temperature (F)
Ambient Pressure (psia)
Ambient Humidity Ratio
Mass Flow Rate (lbm/s)
Low Pressure Compressor
Inlet Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
High Pressure Compressor
Inlet Temperature (F)
Inlet Pressure (psia)
Adiabatic Exit Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Total Gas Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Combustor
Fuel Flow Rate (lbm/s)
Phi
Combustion Effciency
Combustor Exit Temperature (F)
High Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Expansion Ratio
Corrected Flow Rate (lbm/s)
Recuperator
Hot Side dT (F)
Hot Side Heat Rate (Btu/s)
Cold Side dT (F)
Cold Side Heat Rate (Btu/s)
Heat to Ambient (Btu/s)
Effectiveness


2

84.2
14.8
0.0174
0.00

92.5
1.40
17.8
68.0
1.36
0.70
0.72

91.2
20.0
340.7
1.38
150.1
71.5
2.75
1.71
1.29

0.023
0.47
97.1
1496


86.1
14.8
0.0186
0.00

93.0
1.40
14.2
67.3
1.29
0.65
0.67

88.2
19.1
339.3
1.38
153.3
70.6
2.75
1.73
1.37

0.021
0.47
94.4
1399


86.6
14.8
0.0183
0.00

93.9
1.40
14.3
67.9
1.30
0.65
0.67

89.4
19.2
340.5
1.38
152.3
70.3
2.73
1.72
1.36

0.022
0.48
95.2
1435


-92.3
-93.6
197.3
87.5
-6.1
0.32










Run
Tnial
Low Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Efficiency (%)
Expansion Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Hot Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Warm Gas Cooler
Gas Side dT (F)
Water Side dT (F)
Gas Side Heat Rate (Btu/s)
Water Side Heat Rate (Btu/s)
Heat Loss to Ambient (Btu/s)
Effectiveness
Cold Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Overall Performance

Power (hp)
Thermal Efficiency (%)
Average Water Extraction Rate (gph)
Average Fuel Flow Rate (gph)


B4-10
1

923
20.7
1.34
-14.2
63.8
1.39
0.52
0.61

-370.2
-106.6

-417.2
16.4
-115.8
81.4
-34.3
0.68


B4-11
1

963
21.3
1.34
-16.8
63.3
1.44
0.55
0.63

-384.3
-103.0

-442.3
17.7
-113.9
85.2
-28.7
0.69


2

959
21.5
1.34
-17.8
63.8
1.46
0.56
0.63

-382.9
-103.3

-434.2
22.5
-112.7
82.6
-30.1
0.68


930
20.8
1.34
-14.3
63.7
1.39
0.52
0.61

-375.4
-106.8

-417.9
16.1
-114.4
82.6
-31.8
0.68


-64.1 -65.8 -67.2 -69.0
-28.8 -29.2 -29.8 -30.9


1.66
71.7
12.96
0.88
10.98


1.63
71.1
12.56
0.75
11.23


1.46
77.4
13.32
1.02
11.53


1.44
77.5
13.11
1.08
11.73


















































1392 1398 1399 1399
56.5 56.8 57.2 57.7
1.32 1.32 1.32 1.32
-233.5 -231.1 -232.3 -230.6
85.0 85.0 85.0 85.0
2.71 2.66 2.65 2.64
0.85 0.85 0.85 0.84


Table C-5. Data from run B4-12.
Run
Tnial
Main Air Inlet
Ambient Temperature (F)
Ambient Pressure (psia)
Ambient Humidity Ratio
Mass Flow Rate (lbm/s)
Low Pressure Compressor
Inlet Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
High Pressure Compressor
Inlet Temperature (F)
Inlet Pressure (psia)
Adiabatic Exit Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Total Gas Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Combustor
Fuel Flow Rate (lbm/s)
Phi
Combustion Effciency
Combustor Exit Temperature (F)
High Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Expansion Ratio
Corrected Flow Rate (lbm/s)
Recuperator
Hot Side dT (F)
Hot Side Heat Rate (Btu/s)
Cold Side dT (F)
Cold Side Heat Rate (Btu/s)
Heat to Ambient (Btu/s)
Effectiveness


B4-12
1

84.4
14.8
0.0180
0.00

91.3
1.39
22.2
67.6
1.41
0.76
0.78

92.6
20.9
342.8
1.38
157.1
70.3
2.71
1.78
1.30

0.023
0.43
95.8
1448


2 3 4


85.7
14.8
0.0183
0.00

93.7
1.39
24.1
68.7
1.45
0.78
0.80

99.5
21.4
350.1
1.38
157.9
69.5
2.66
1.79
1.28

0.023
0.42
96.4
1457


86.5
14.8
0.0186
0.00

95.0
1.39
24.9
69.1
1.46
0.79
0.81

99.8
21.6
349.2
1.38
158.0
69.7
2.65
1.80
1.27

0.023
0.41
96.5
1460


87.1
14.8
0.0186
0.00

95.8
1.39
26.3
69.3
1.48
0.80
0.82

101.9
21.9
351.6
1.38
157.8
69.6
2.64
1.79
1.25

0.023
0.41
96.5
1473


-98.1
-98.8
202.8
92.5
-6.2
0.32


-105.1
-97.9
200.5
91.8
-6.1
0.32


-102.1
-98.6
201.0
92.6
-6.0
0.32


-97.6
-98.6
201.9
92.7
-5.9
0.31










Run
Tnial
Low Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Efficiency (%)
Expansion Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Hot Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Warm Gas Cooler
Gas Side dT (F)
Water Side dT (F)
Gas Side Heat Rate (Btu/s)
Water Side Heat Rate (Btu/s)
Heat Loss to Ambient (Btu/s)
Effectiveness
Cold Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Overall Performance

Power (hp)
Thermal Efficiency (%)
Average Water Extraction Rate (gph)
Average Fuel Flow Rate (gph)


B4-12
1

954
22.6
1.34
-22.2
65.2
1.53
0.61
0.66

-382.2
-104.6

-408.1
17.7
-108.1
81.7
-26.3
0.65


2 3 4


957
23.1
1.34
-24.1
65.6
1.56
0.63
0.67

-392.4
-105.9

-403.2
18.4
-105.0
83.4
-21.6
0.64


962
23.3
1.34
-24.9
65.8
1.57
0.64
0.67

-389.2
-105.2

-404.4
18.7
-105.5
84.7
-20.8
0.64


968
23.5
1.34
-26.3
66.5
1.59
0.65
0.67

-394.8
-104.8

-402.2
18.9
-103.2
85.8
-17.4
0.64


-82.4 -78.9 -83.4 -84.3
-37.8 -36.1 -38.3 -38.7


1.35
76.4
13.05
1.07
11.62


1.30
73.2
12.56
0.85
11.56


1.29
74.3
12.66
0.85
11.65


1.25
72.9
12.27
0.90
11.78


















































1373 1388 1383 1367
45.4 45.3 50.1 52.5
1.32 1.32 1.32 1.32
-192.7 -191.0 -214.3 -217.8
85.0 85.0 85.0 85.0
2.90 2.90 2.83 2.74
0.85 0.84 0.86 0.85


Table C-6. Data from run B4-14.
Run
Tnial
Main Air Inlet
Ambient Temperature (F)
Ambient Pressure (psia)
Ambient Humidity Ratio
Mass Flow Rate (lbm/s)
Low Pressure Compressor
Inlet Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
High Pressure Compressor
Inlet Temperature (F)
Inlet Pressure (psia)
Adiabatic Exit Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Total Gas Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Combustor
Fuel Flow Rate (lbm/s)
Phi
Combustion Effciency
Combustor Exit Temperature (F)
High Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Expansion Ratio
Corrected Flow Rate (lbm/s)
Recuperator
Hot Side dT (F)
Hot Side Heat Rate (Btu/s)
Cold Side dT (F)
Cold Side Heat Rate (Btu/s)
Heat to Ambient (Btu/s)
Effectiveness


B4-14
2

84.3
14.7
0.0171
0.12

89.8
1.40
5.1
34.8
1.06
0.55
0.57

72.3
15.6
327.4
1.39
125.7
72.1
2.90
1.41
1.34

0.019
0.40
99.7
1503


3 4 5


85.8
14.7
0.0176
0.11

92.9
1.40
5.3
33.2
1.06
0.55
0.57

74.0
15.6
329.2
1.39
123.7
72.3
2.90
1.38
1.32

0.019
0.40
99.9
1528


86.8
14.7
0.0183
0.00

94.5
1.40
12.2
55.9
1.20
0.66
0.68

81.5
17.7
332.7
1.38
140.1
72.3
2.83
1.59
1.35

0.021
0.45
97.6
1473


88.6
14.7
0.0181
0.00

97.5
1.39
14.5
66.2
1.30
0.65
0.67

91.0
19.1
342.9
1.38
148.0
70.6
2.74
1.67
1.32

0.021
0.47
95.1
1425


-127.0
-83.4
215.5
77.0
-6.4
0.30


-112.5
-83.4
219.6
77.3
-6.1
0.30


-85.0
-92.5
213.8
86.5
-6.0
0.31


-101.5
-89.6
196.3
83.9
-5.8
0.32










Run
Tnial
Low Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Efficiency (%)
Expansion Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Hot Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Warm Gas Cooler
Gas Side dT (F)
Water Side dT (F)
Gas Side Heat Rate (Btu/s)
Water Side Heat Rate (Btu/s)
Heat Loss to Ambient (Btu/s)
Effectiveness
Cold Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Overall Performance

Power (hp)
Thermal Efficiency (%)
Average Water Extraction Rate (gph)
Average Fuel Flow Rate (gph)


B4-14
2

889
17.9
1.34
-5.1
-3.2
1.21
0.39
0.52

-411.1
-81.0

-390.1
11.7
-76.6
55.6
-21.0
0.70


3 4 5


916
17.9
1.34
-5.3
25.7
1.22
0.39
0.52

-417.3
-79.7

-391.7
12.6
-73.8
59.6
-14.2
0.68


947
20.1
1.34
-12.2
59.8
1.37
0.50
0.60

-404.0
-99.7

-420.2
15.6
-101.0
72.9
-28.1
0.68


925
20.6
1.34
-14.5
63.4
1.41
0.53
0.61

-374.5
-102.0

-398.2
17.5
-104.7
81.3
-23.4
0.64


-52.6 -60.5 -67.1 -80.8
-20.0 -21.4 -28.5 -34.4


1.11
67.0
13.86
2.64
9.59


1.08
67.3
13.88
1.94
9.62


1.39
74.2
13.85
1.71
10.64


1.58
69.7
12.86
0.77
10.76



























































-96.3
-94.1
207.2
88.0
-6.1
0.32


-100.8
-100.4
204.7
94.4
-6.0
0.33


-97.3
-83.8
215.0
77.5
-6.2
0.30


1382 1403 1406 1372
45.2 52.5 56.3 45.1
1.32 1.32 1.32 1.32
-191.7 -222.4 -238.5 -193.4
85.0 85.0 85.0 85.0
2.91 2.76 2.71 2.89
0.85 0.85 0.86 0.86


Table C-7. Data from runs B4-15 and B4-16
Run B4-15


1


B4-16
2 3


Tnial
Main Air Inlet
Ambient Temperature (F)
Ambient Pressure (psia)
Ambient Humidity Ratio
Mass Flow Rate (lbm/s)
Low Pressure Compressor
Inlet Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
High Pressure Compressor
Inlet Temperature (F)
Inlet Pressure (psia)
Adiabatic Exit Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Total Gas Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Combustor
Fuel Flow Rate (lbm/s)
Phi
Combustion Effciency
Combustor Exit Temperature (F)
High Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Expansion Ratio
Corrected Flow Rate (lbm/s)
Recuperator
Hot Side dT (F)
Hot Side Heat Rate (Btu/s)
Cold Side dT (F)
Cold Side Heat Rate (Btu/s)
Heat to Ambient (Btu/s)
Effectiveness


1

85.1
14.7
0.0186
0.13

91.3
1.40
5.0
36.7
1.06
0.55
0.56

72.6
15.6
327.6
1.39
126.7
71.8
2.89
1.42
1.35

0.019
0.41
99.6
1510


83.0
14.7
0.0166
0.12

88.3
1.40
5.2
30.7
1.06
0.54
0.56

70.9
15.5
326.7
1.39
124.4
71.9
2.91
1.39
1.33

0.019
0.41
99.6
1520


86.1
14.7
0.0171
0.00

93.9
1.40
16.5
61.6
1.29
0.70
0.72

85.9
19.0
337.3
1.38
146.9
70.7
2.76
1.66
1.32

0.021
0.43
96.3
1456


86.5
14.7
0.0171
0.00

94.8
1.39
24.0
64.6
1.41
0.79
0.81

91.0
20.7
340.7
1.38
158.4
70.5
2.71
1.80
1.32

0.022
0.41
96.5
1438


-113.3
-82.2
214.5
75.8
-6.5
0.30















































V \VI j


Run
Tnial
Low Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Efficiency (%)
Expansion Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Hot Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Warm Gas Cooler
Gas Side dT (F)
Water Side dT (F)
Gas Side Heat Rate (Btu/s)
Water Side Heat Rate (Btu/s)
Heat Loss to Ambient (Btu/s)
Effectiveness
Cold Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Overall Performance

Power (hp)
Thermal Efficiency (%)
Average Water Extraction Rate (gph)
Average Fuel Flow Rate (gPh)


B4-15
1


B4-16
2 3


1

919
17.9
1.34
-5.0
44.3
1.22
0.39
0.52


910
17.7
1.34
-5.2
55.1
1.21
0.38
0.52


959
21.0
1.34
-16.5
64.5
1.43
0.54
0.62


961
22.8
1.34
-24.0
66.7
1.55
0.62
0.67


-410.0 -390.4 -385.6 -415.0
-79.8 -100.2 -104.7 -83.0


-396.1
10.7
-77.0
59.0
-18.0
0.70

-57.1
-21.4

1.11
67.2
13.82
2.78
9.65


-407.0
16.2
-100.9
80.1
-20.8
0.64

-89.0
-37.9

1.37
75.5
13.81
0.96
10.85


-406.6
17.9
-106.5
84.7
-21.8
0.64

-93.9
-43.1

1.29
80.2
13.89
0.85
11.45


-399.2
10.3
-80.8
61.6
-19.3
0.70

-58.2
-21.7

1.12
66.7
13.60
3.16
9.73

















































1377 1415
48.0 53.9
1.32 1.32
-204.8 -231.5
85.0 85.0
2.87 2.76
0.85 0.86


Table C-8. Data from run B4-17.
Run
Tnial
Main Air Inlet
Ambient Temperature (F)
Ambient Pressure (psia)
Ambient Humidity Ratio
Mass Flow Rate (lbm/s)
Low Pressure Compressor
Inlet Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
High Pressure Compressor
Inlet Temperature (F)
Inlet Pressure (psia)
Adiabatic Exit Temperature (F)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Pressure Ratio
Total Gas Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Combustor
Fuel Flow Rate (lbm/s)
Phi
Combustion Effciency
Combustor Exit Temperature (F)
High Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Effciency (%)
Expansion Ratio
Corrected Flow Rate (lbm/s)
Recuperator
Hot Side dT (F)
Hot Side Heat Rate (Btu/s)
Cold Side dT (F)
Cold Side Heat Rate (Btu/s)
Heat to Ambient (Btu/s)
Effectiveness


B4-17
1

86.2
14.7
0.0177
0.00

93.1
1.40
9.2
47.6
1.14
0.62
0.64

78.1
16.7
330.4
1.38
133.4
72.6
2.87
1.51
1.35

0.020
0.46
95.7
1463


2

89.2
14.7
0.0180
0.00

96.3
1.39
19.9
61.2
1.33
0.75
0.78

84.8
19.6
334.9
1.38
150.9
70.9
2.76
1.72
1.32

0.021
0.41
96.2
1444


-73.6
-89.8
218.3
83.7
-6.1
0.32


-101.6
-98.0
210.0
92.1
-5.9
0.33














































V \V j


Run
Tnial
Low Pressure Turbine
Inlet Temperature (F)
Inlet Pressure (psia)
Gamma
Shaft Power (hp)
Isentropic Efficiency (%)
Expansion Ratio
Flow Rate (lbm/s)
Corrected Flow Rate (lbm/s)
Hot Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Warm Gas Cooler
Gas Side dT (F)
Water Side dT (F)
Gas Side Heat Rate (Btu/s)
Water Side Heat Rate (Btu/s)
Heat Loss to Ambient (Btu/s)
Effectiveness
Cold Gas Cooler
Gas Side dT (F)
Gas Side Heat Rate (Btu/s)
Overall Performance

Power (hp)
Thermal Efficiency (%)
Average Water Extraction Rate (gph)
Average Fuel Flow Rate (gPh)


B4-17
1


2

963
21.9
1.34
-19.9
65.2
1.49
0.58
0.65


950
19.1
1.34
-9.2
59.2
1.30
0.45
0.57


-405.6 -401.0
-96.0 -103.7


-433.1
14.1
-100.8
76.2
-24.6
0.71

-58.6
-23.9

1.42
71.4
13.84
2.15
10.24


-429.5
15.2
-108.0
81.4
-26.6
0.68

-79.6
-35.3

1.29
80.6
14.49
1.40
11.04










LIST OF REFERENCES


[1] Gasparovic, N., 1968, "The Advantage of Semi-Closed Cycle Gas Turbines for
Naval Ship Propulsion," Naval Engineers Journal 80, April 1968, pp. 275-281,
333

[2] Nemec, T. S., 1995, "Thermodynamic Design Point Study of a Semi-Closed
Recuperated Intercooled Gas Turbine Combined with a Rankine Bottoming
Cycle," Master's Thesis, Dept. of Mechanical Engineering, University of Florida.

[3] Landon, J. C., 1996, "Design and Off-Design Point Study of Two Regenerative
Feedback Gas Turbine Engines for Marine Applications," Master' s Thesis, Dept.
of Mechanical Engineering, University of Florida.

[4] Danias, G. E. 1998, "Design and Off-Design Point Study of Two Regenerative
Feedback Gas Turbine Engines for Helicopter Applications," Master's Thesis,
Dept. of Mechanical Engineering, University of Florida.

[5] MacFarlane, R. S., 1997, "A Study of the Impact of Water Extraction on the
Regenerative Feedback Turbine Engine Cycle," Master' s Thesis, Dept. of
Mechanical Engineering, University of Florida.

[6] Muley, N. S., 2002, "Effect of Exhaust Gas Recirculation on Thermal NOx
Formation Rate in a Gas Turbine Engine," Master' s Thesis, Dept. of Mechanical
Engineering, University of Florida.

[7] Boza, J. J., 2003, "Performance of a Semi-Closed Gas Turbine and Absorption
Refrigeration Combined Cycle," Master's Thesis, Dept. of Mechanical
Engineering, University of Florida.

[8] Khan, J., 2006, "Design and Optimization of a Distributed Generation System with
the Production of Water and Refrigeration," Doctoral Dissertation, Dept. of
Mechanical Engineering, University of Florida.

[9] Lear, W. E., Laganelli, A. L., 1999, "High Pressure Regenerative Turbine Engine:
21st Century Propulsion," Final Test Report for Contract No. NAS3-27396.

[10] Crittenden, J. F., Lear, W. E., Azzazy, M., 1999, "Exploratory Design of a Depleted
Oxygen Gas Turbine Combustor," Final Report for Contract No. NAS3-27759.

[1l] Pringle, D. S., 1972, "Testing a Small Gas Turbine In the Laboratory Shop,"
Saw/yer 's Gas Turbine Engineering Handbook, Vol. III, Gas Turbine Publications,
Inc.

[12] Lear, W. E., 2006, "Test Report for a Novel Combined Cycle Turbine Engine with
Water Harvesting," Test Report for Subcontract to Phase II SBIR Contract No.
AO3-037.










[13] Brown, D., 2005, Internal Report for US Army Phase II SBIR Contract No. AO3-
037.

[ 14] Gater, R. A., 2002, Engmneering Thermodynamics-Foundation Topics, Textbook
Published by Author.

[ 15] Incropera, F. P., DeWitt, D. P., 2002, Fundamentalsd~~~dd~~~ddd~~ of Heat and M~a;ss Transfer, 5th
Edition, John Wiley & Sons, New York.

[16] 1958, Rover Gas Turbine-Service Literature, Rover Gas Turbines Ltd., Solihull
Warwickshire England, Publication No. SP/101/758.

[17] Crittenden, J. F., 1999, "Dilute, Kinetically Controlled Combustion Efficiency
Prediction for Recirculating Semi-Closed Gas Turbine Cycles: A Non-
Dimensional Approach Using the First Damkohler Number as a Parameter,"
Master' s Thesis, Dept. of Mechanical Engineering, University of Florida, pp. 56-
57.

[18] Holman, J. P. 2001, Experimental M'ethods for Engmneers, 7th Edition, McGraw-
Hill, Boston.

[19] Volponi, A. J., 1999, "Gas Turbine Parameter Corrections," Journal of Engineering
for Gas Turbines and Power, 121 Oct. pp. 613-621.

[20] Brown, D. Internal Progress Report for US Army Phase II SBIR Contract No. AO3-
037.









BIOGRAPHICAL SKETCH

Eric Howell received his Bachelor of Science degree cum laude in mechanical engineering

from the University of Florida in 2004. His deep-seated fascination with the thermal sciences

compelled him to pursue a master' s degree at the University of Florida, while working as a

Research Assistant in the Energy and Gasdynamics Systems Laboratory. The Author received

his Master of Science degree in mechanical engineering from the University of Florida in 2007.





PAGE 1

1 EXPERIMENTAL STUDY OF A NOVEL GAS TURBINE ENGINE INTEGRATED WITH AN ABSORPTION REFRIGERATION SYSTEM by ERIC B. HOWELL A THESIS PRESENTED TO THE GRADUATE SCHOOL OF THE UNIVERSITY OF FLOR IDA IN PARTIAL FULFILLMENT OF THE REQUIREMENTS FOR THE DEGREE OF MASTER OF SCIENCE UNIVERSITY OF FLORIDA 2007

PAGE 2

2 2007 Eric B. Howell

PAGE 3

3 ACKNOWLEDGMENTS I first thank my committee chair, Dr. William L ear, for giving me the great opportunity to be a part of this research effort, and for his c ounsel throughout my graduate studies. I would also thank Dr. S. A. Sherif, and Dr. Herbert Ingley for serving on my committee and demonstrating great patience with me throughout this work. I must also thank Mr. John Crittenden for all of his advice, encouragement, and comradeship during this undertaking. I am also indebted to Mr. Dan Brown and Mr. Todd Nemec for all their technical assistance and recommendations.

PAGE 4

4 TABLE OF CONTENTS page ACKNOWLEDGMENTS...............................................................................................................3 LIST OF TABLES................................................................................................................. ..........7 LIST OF FIGURES................................................................................................................ .........8 LIST OF TERMS.................................................................................................................. .........10 ABSTRACT....................................................................................................................... ............14 CHAPTER 1 INTRODUCTION..................................................................................................................16 2 LITERATURE REVIEW.......................................................................................................21 History of Semi-Closed Cycle Concepts................................................................................21 Analytical Studies............................................................................................................. ......22 Experimental Studies........................................................................................................... ...24 3 TESTING HARDWARE........................................................................................................26 Combined Cycle Components................................................................................................26 Engine Characteristics and Specifications.......................................................................26 Recuperator.................................................................................................................... ..27 Turbocharger...................................................................................................................27 Heat Exchangers..............................................................................................................27 Hot gas cooler (HGC)..............................................................................................28 Warm gas cooler (WGC)..........................................................................................28 Cold gas cooler (CGC).............................................................................................28 Dynamometer..................................................................................................................29 Ducting........................................................................................................................ ....29 Facility Resources............................................................................................................. ......29 Chilled Water.................................................................................................................. .29 Fans and HVAC..............................................................................................................30 Vapor Absorption Refrigeration System (VARS)..................................................................30 Overview of Build Configurations.........................................................................................30 Build 1........................................................................................................................ .....31 Build 2A....................................................................................................................... ...32 Build 3........................................................................................................................ .....33 Build 4........................................................................................................................ .....34

PAGE 5

5 Data Measurement............................................................................................................... ...34 Data Acquisition Hardware.............................................................................................34 Data Acquisition Software..............................................................................................36 Instrumentation................................................................................................................36 Thermocouples.........................................................................................................36 Pressure transducers.................................................................................................37 Turbine and paddle-wheel flowmeters.....................................................................37 Optical tachometer...................................................................................................37 Load cells.................................................................................................................38 Gas analysis..............................................................................................................38 Mechanical instruments and analog readers.............................................................39 4 TEST PROCEDURES............................................................................................................45 Test Plan...................................................................................................................... ...........45 Engine/Component Preparation..............................................................................................45 Instrument Preparation......................................................................................................... ...46 Transducer Offsets...........................................................................................................46 Ambient Conditions.........................................................................................................47 5 DATA REDUCTION SCHEME............................................................................................49 Gas Properties................................................................................................................. ........49 Enthalpy and Specific Heat.............................................................................................50 Gas Composition.............................................................................................................52 Basic Equations and Assumptions..........................................................................................54 Main Air Inlet (MAI).......................................................................................................54 Low Pressure Compressor (LPC)....................................................................................55 Recirculation Venturi (RCV)..........................................................................................55 Hot Gas Cooler (HGC)....................................................................................................56 Warm Gas Cooler (WGC)...............................................................................................56 Cold Gas Cooler (CGC)..................................................................................................58 Ducting Section 4-5.........................................................................................................59 High Pressure Compressor (HPC)...................................................................................60 Combustor...................................................................................................................... .61 Ducting Section 7-8.........................................................................................................63 High Pressure Turbine (HPT)..........................................................................................64 Recuperator.................................................................................................................... ..66 Low Pressure Turbine (LPT)...........................................................................................68 Mixing Junctions.............................................................................................................68 Propagation of Uncertainty.....................................................................................................70 Data Adjustments............................................................................................................... .....72 Pressure Drop Considerations.................................................................................................73 6 EXPERIMENTAL RESULTS...............................................................................................79

PAGE 6

6 7 CONCLUSIONS AND RECOMMENDATIONS...............................................................105 Summary........................................................................................................................ .......105 Recommendations................................................................................................................ .107 APPENDIX A OPERATING PROCEDURES FOR HPRTE/VARS..........................................................110 B SETUP PROCEDURES FOR HPRTE/VARS.....................................................................113 C ADJUSTED EXPERIMENTAL DATA..............................................................................119 LIST OF REFERENCES.............................................................................................................135 BIOGRAPHICAL SKETCH.......................................................................................................137

PAGE 7

7 LIST OF TABLES Table page 3-1. Temperature locations and instrumentation..........................................................................40 3-2. Pressure locations and instrumentation.................................................................................41 5-1. Uncertainty of instruments............................................................................................... .....74 5-2. Data scaling parameters [19]............................................................................................. ....75 C-1. Data from runs B4-1 and B4-2............................................................................................119 C-2. Data from runs B4-3 and B4-5............................................................................................121 C-3. Data from runs B4-7 and B4-8............................................................................................123 C-4. Data from runs B4-10 and B4-11........................................................................................125 C-5. Data from run B4-12...................................................................................................... .....127 C-6. Data from run B4-14...................................................................................................... .....129 C-7. Data from runs B4-15 and B4-16........................................................................................131 C-8. Data from run B4-17...................................................................................................... .....133

PAGE 8

8 LIST OF FIGURES Figure page 1-1. Block diagram of HPRTE flow path.....................................................................................20 3-1. Block diagram of Build 1 configuration................................................................................42 3-2. Photograph of Build 1 improvi sed evaporator installation....................................................43 3-3. Block diagram of Bu ild 2A configuration.............................................................................43 3-4. Block diagram of Build 3 configuration................................................................................44 3-5. Block diagram of Build 4 configuration................................................................................44 5-1. Map used for finding the HPC adiabatic exit temperature, shown as a function of corrected flow rate and corrected speed [16].....................................................................76 5-2. Rover 1S-60 internal power losses as a function of rotor speed [16]....................................77 6-1. Actual net power output from Build 4 engine runs...............................................................92 6-2. Corrected Power, power, and ther mal efficiency over corrected speed................................93 6-3. Scaling parameters versus corrected speed...........................................................................93 6-4. Theta versus LPC pressure ratio.......................................................................................... ..94 6-5. Power versus LPC pressure ratio.......................................................................................... .94 6-6. High pressure turbomachinery temperature differences and mass flow rate versus LPC pressure ratio................................................................................................................. .....95 6-7. Percent change in high pr essure turbomachinery temperature differences and mass flow rate versus LPC pressure ratio...........................................................................................95 6-8. Flow parameters versus corrected speed...............................................................................96 6-9. Generalized effects of HPC pressure ratio and CTR on specific power, illustrated on a T-S diagram.................................................................................................................... ...96 6-10. Specific power versus CTR, fo r various HPC pressure ratios.............................................97 6-11. Corrected power versus CTR fo r various HPC pressure ratios...........................................97 6-12. HPC pressure ratio and CTR versus corrected speed..........................................................98 6-13. Thermal efficiency versus CTR...........................................................................................98

PAGE 9

9 6-14. Thermal efficiency ve rsus HPC pressure ratio....................................................................99 6-15. HPC isentropic efficien cy versus corrected speed..............................................................99 6-16. Recirculation ratio ve rsus LPC pressure ratio...................................................................100 6-17. Fresh air flow rate ve rsus recirculation ratio.....................................................................100 6-18. Equivalence ratio vers us recirculation ratio......................................................................101 6-19. Equivalence ratio vers us LPC pressure ratio.....................................................................101 6-20. Nitric oxide and carbon monoxide con centrations versus r ecirculation ratio...................102 6-21. Oxygen and nitrogen concentrat ions versus recirculation ratio........................................102 6-22. Water-to-fuel ratio ve rsus LPC pressure ratio...................................................................103 6-23. Recirculation flow rate versus LPC pressure ratio............................................................103 6-24. Temperatures related to water ex traction versus LPC pressure ratio................................104

PAGE 10

10 LIST OF TERMS Roman Letters a first degree coefficient for specific heat polynomial A cross sectional area b second degree coefficient for specific heat polynomial c third degree coefficient for specific heat polynomial cP constant-pressure specific heat e heat exchanger effectiveness F force FA fuel-to-air ratio h specific enthalpy or convection coefficient M molecular weight m mass flow rate n moles n mole flow rate N engine speed N* corrected engine speed P pressure or power P* corrected power PR pressure ratio Q heat transfer rate R mass-specific gas constant or recirculation ratio R mole-specific (universal) gas constant

PAGE 11

11 T temperature V volumetric flow rate W power y mole fraction Greek Letters specific heat ratio pressure correction parameter emissivity efficiency temperature correction parameter density Stefan-Boltzmann constant dimensionless temperature parameter angular velocity, or uncertainty Subscripts 0 reference state, 300 Kelvin AMB ambient AVG average COLD cold side of heat exchanger COMB combustor CONV convective DAQ data acquisition DRY dry gas mixture, omitting water

PAGE 12

12 DYNO dynamometer fg liquid-vapor transition HOT hot side of heat exchanger HPTI high pressure turbine inlet HPTX high pressure turbine exit i ith term in a series, or ith constituent of a mixture in general inlet state point INST instrumentation LPCI low pressure compressor inlet LPCX low pressure compressor exit max maximum possible value MIX mixture MECH mechanical friction NC non-condensable out general outlet state point R recuperator RAD radiative RCVI recirculation venturi inlet SEN sensible SURF surface TOT total value for mixture TRAN transient W water

PAGE 13

13 WGCG warm gas cooler, gas side WGCI warm gas cooler inlet WGCW warm gas cooler, water side X generalized independent parameter

PAGE 14

14 Abstract of Thesis Presen ted to the Graduate School of the University of Florida in Partial Fulfillment of the Requirements for the Degree of Master of Science EXPERIMENTAL STUDY OF A NOVEL GAS TURBINE ENGINE INTEGRATED WITH AN ABSORPTION REFRIGERATION SYSTEM by Eric B. Howell August 2007 Chair: William E. Lear Major: Mechanical Engineering A series of tests were performed on a nove l gas turbine engine, coupled to a vapor absorption refrigeration system. The engine cy cle is unique in that it is semi-closed and turbocharged in addition to uti lizing a recuperator. The refrig eration system is in place to remove heat from the recirculating exhaust gases, to cool the engine inlet gases, and to produce additional refrigeration for auxiliary thermal loads. The engine was developed in several stages and operated over a wide range of conditions. Th e data from these experiments were reduced, analyzed and presented with emphasis placed on th e unique attributes of the combined cycle. Observed hardware limitations are also identified and adjustments were made to the data in an effort to estimate certain performance pa rameters for more ideal components. It was observed in these experiments that power transitions could be made with this cycle by varying the turbocharger pressu re ratio. Doing so results in a somewhat constant thermal efficiency over a range of pow er levels, since the core engine approximately remains at a constant, dimensionless operating po int. It is also shown that semi-closure of the power cycle reduces the air flow necessary to run the engine, highlighting the anticipated size reduction of certain components. The data also validate that semi-closure of the cycle serves to alter the combustion environment, such that NOX emissions are reduced, making the power cycle more

PAGE 15

15 attractive from an environmental perspective. The cooling of recirculation gases presents the additional opportunity to extract fresh water from th e gas path. This capability is also realized in these experiments, but the amount of water extr acted was limited by some of the experimental hardware to less than half the theoretical maximum.

PAGE 16

16 CHAPTER 1 INTRODUCTION With rising energy costs and instability of energy sources, it is becoming ever more important to develop newer technologies to meet the unrelenting increase in demand for energy. This challenge is further complicated by concerns about the current and future states of the environment, and must be overcome while concu rrently decreasing any negative, environmental impacts. Distributed generation of energy is seen as a potential solution, at least in part, to present energy challenges. Some varieties of distributed generati on exist in the form of solar panels and wind turbines, which use natural en ergy sources that are often inconsistent and dependant upon weather. Moreover, these type s of systems must be implemented on a large scale if they are to be signif icant contributors to the energy supply. Microturbines are also becoming more attractive as distributed generatio n systems, as they are more compact and can burn a large variety of fuels. Microturbines also pr ovide the capability of supporting heating loads with exhaust gases, and mo re recent innovations utilize abso rption refrigeration systems to provide refrigeration loads as well. It is hypothesized that seve ral innovations to the conventio nal gas turbine engine can increase its usefulness as a distributed gene ration system. The particular improvements implemented in this study are be lieved to reduce the phys ical size of the engine and enable it to follow load demands without signifi cant sacrifices of thermal e fficiency. In addition, it is theoretically possible for the integration of such an engine with a Vapor Absorption Refrigeration System (VARS) to further enha nce its attractiveness by enabling the combined cycle to produce external refr igeration, and fresh water. In an effort to better quantify the extent of these prospective capab ilities, modeling and experimental activities were conducted at the Univ ersity of Florida. This thesis details the

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17 experimental efforts undertaken on a novel gas tu rbine engine cycle combined with a VARS. The power cycle is unique in that it is both se mi-closed and turbocharge d, in addition to being recuperated, and is referred to as the High Pressure Regenera tive Turbine Engine (HPRTE). Figure 1-1 shows the flow path of the HPRTE. There are two main reasons for turbocharging the gas turbine engine. The first, more apparent of these is the resulting increase in pow er density. By increasin g the pressure of the working fluid, the density is also increased. In turn, the same size engi ne can operate with a greatly increased mass flow rate which equates to more power from a smaller engine. The second reason for turbocharging th e gas turbine engine is more subtle. The engine thermal efficiency can remain approximately constant over a range of power levels. This is accomplished by using the turbocharger itself to perform power transitions. Exhaust can be diverted through/around the Low Pressure Turbine (L PT) to raise/lower the inlet pressure of the engine. If the inlet temperature is held constant then the density and flow rate will in turn increase proportionately to the inlet pressure. Do ing so allows the core en gine to remain at the same, dimensionless operating point (same corrected flow corrected power and pressure ratios) for different power levels. The core engine is simply the recuperated gas turbine engine being operated apart from other, supporting hard ware (turbocharger, VARS, etc.). It is also theorized that the semi-closure of the gas turbine cycle also offers several advantages over conventional gas turb ine cycles. The first of these relates to the fact that about 20 percent of the air ingested by a conventional gas turbine engine is necessary to support combustion. Semi-closing the cycle entails recircul ating much of the exhaust gases to be used again as the working fluid, and as this recirculated proportion of gas is increased, the net inflow and outflow of air and exhaust gas are reduced. Th is effect is believed to significantly reduce the

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18 size of the turbocharger and othe r inlet air and exhaus t ducting. Some applications may also find significant utility in the smaller thermal and acous tic signatures from the engine resulting from the reduced exhaust flow. Speculations are also made regarding the at tractiveness of the semi-closed gas turbine engine from an environmental perspective. Th e more oxygen-dilute gases are expected to reduce the temperature within the primar y zone of the combustor. This should result in a significant reduction of oxides of nitrogen (NOX), since NOX formation rates scale very strongly with temperature. Similar arguments can also be made about soot formation rates. Recirculating significant quanti ties of hot exhaust gases to the High Pressure Compressor (HPC) inlet presents both the pr oblem of significant cooling, a nd the benefit of fresh water extraction. Aside from its front -end material considerations, the gas turbine engine would perform poorly with such extreme temperatures of mixed inlet gase s. Hence, these gases must be cooled. However, a benefit is made available by the cooling of these gases, as an appreciable fraction of the recirculation flow is water vapo r. Once condensed in the cooling process, the water can straightforwardly be ex tracted and made ready to dri nk. This rather large cooling challenge presented by the recircul ating exhaust gases can be viewed as an opportunity to extract further potential from the power cycl e via integration of the VARS. The VARS is driven by the heat rejected from the recirculation loop and in turn benefits the HPRTE in several ways. The most notable of these is the theoretical capability it offers for reducing the HPC inlet temperature to below am bient temperatures, and improving the engine performance. This attribute im plies that the engine performance should depend much less, if at all, on ambient conditions, since the HPC inlet temp erature and pressure can be controlled by the VARS and turbocharger, respectively. Additiona lly, it is theorized that the VARS can produce

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19 more refrigeration than is necessary for the gaspath cooling. As a result, a fourth commodity, refrigeration, can be counted alongside of pow er, heat, and water as an output from this combined cycle. The apparent advantages and capabilities of this combined cycle motivated the US Army Research Laboratory to fund the development of an experimental HPRTE/VARS combined cycle for the purpose of demonstrating its therm odynamic principles. The University of Florida was subcontracted to carry out the design and test ing of several experimental apparatuses within its Energy and Gasdynamics Laboratory (EGDL). The overall objectives of this research work were to assemble an experimental HPRTE/VARS combined cycle, demonstrate its operability, and demonstrate that its theore tical capabilities could be re alized in accordance with the predictions of thermodynamic models. A stepwi se approach was taken in the experimental development such that the capabilities of the experimental apparatus were expanded incrementally. A second objective of the experime ntal work was to help develop and validate design-point models for this combined cycle. Though details of the modeling efforts are not within the scope of this work, some comp arisons are made between the modeled and experimental results.

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20 Figure 1-1. Block diagram of HPRTE flow path.

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21 CHAPTER 2 LITERATURE REVIEW History of Semi-Closed Cycle Concepts Innovations in the Brayton cycle have continue d to improve the perfor mance of gas turbine engines since their inception. One such improve, a nd perhaps the most pertin ent to this work is the semi-closure of the Brayton cycle. Alt hough little development ha s been done recently, the concept of semi-closure is nearly as old as ga s turbines, as summarized by Gasparovic [1]. As early as 1940s, attempts were made at developing semi-closed gas turbine engines. The initial attempts at recirculating exhaust gases in a semi-closed cycle were plagued with corrosion associated with the heavy fuel oils of the time Subsequent efforts only recirculated air in a secondary, charging circuit that added excessi ve complications and costs to the cycle. In the late 1960s Gasparovic attempted to re vive the previously abandoned idea of semicloser, as fuel technology ha d progressed. He also presented the idea of semi-closure implemented with a low-pressure turbocharging system. In [1], he notes the potential of gas turbine engines of a given power to be made smaller, estimating a pound-per-horsepower of about 2/3 that of conven tional engines. He also points out th e advantage of part-load efficiency achieved by varying the low-pressure charging of the cycle. Specifically, he cites the Sulzer Brothers design for a 20 MW plant operating at 32% efficiency at full power, and at 28% efficiency at only half power. Again, this engi ne eventually failed due to excessive corrosion. In recent years several research efforts at the University of Fl orida have once again revitalized the semi-closed, turbocharged, gas turb ine engine referred to as the HPRTE. These studies have focused on many of the specifi c HPRTE attributes, as well as on various applications thought to be well suited by this cy cle. However, most of the more recent, University of Florida studies were modeling analyses.

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22 Analytical Studies Nemec [2] conducted studies on the HPRTE with a Rankine bottoming cycle. In his work, several bottoming cycle fluids we re considered, as well as feed water heating, superheating and recuperation in relation to optimizing the overall thermal efficiency. One apparent challenge in Nemecs analysis was matching the bottoming-cycle performance with that of the topping cycle. However, with the fluids and components cons idered, an optimal ther mal efficiency of 54.5% was reported. Nemec notes that higher efficienci es are possible with the inclusion of different bottoming fluids and combined cycle components. Landon [3] and Danias [4] performed several in dependent, but similar analytical studies on the HPRTE. Landons analysis evaluated the suit ability of several HPRTE configurations for marine applications and compared them with currently employed technology. Those results indicate significantly higher pa rt-load efficiencies, resulting in a 24% increase in range for a given marine vessel. Danias analysis also re vealed similar qualities from two HPRTE cycles, but for helicopter applications. His model predicted an increase in non-dimensional range of up to 46% over a conventional helicopter engine resulting from the improved efficiency at offdesign power levels. Danias simulations also pr edicted design-point effi ciency increases of 3035% over the baseline engine, at the expense of a bearing a slightly larger and heavier recuperated engine. MacFarlane [5] investigated the implications of extracting water from the HPRTE recirculation flow path. He also looked into a ny potential benefits of wa ter re-injection into the cycle. His results show a ma ximum possible decrease of Specif ic Fuel Consumption (SFC) of about 8%. Furthermore, MacFarla ne found that water extraction c ould also increase the specific power by about 4.5% over baseline. Re-injection of water back into the cycle seemed to increase SFC by about 2%, with little or no other impact on other performance parameters. However, it is

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23 noted that the increase in SFC come s with the injection of water in its liquid phase, and that this increase would be less if the wa ter were injected as steam. Muley [6] carried out studies on the emissions-related effects of semi-closure. In his work, Muley focused primarily on how the formation of thermal NOX was impacted by recirculating exhaust gases. He discovered that an impressive reduction of thermal NOX could be realized by semi-closure of a gas turbine power cycle. This effect grew more pronounced with further dilution of the combustor inlet gases, yielding a decrease in thermal NOX of about eight to nine orders of magnitude, depending also on the combustor inlet temperature. Boza [7] was the first to perform a mode ling analysis on the HPRTE power cycle combined with a VARS in 2003. In his mode ling efforts, Boza uncovered some striking performance characteristics made possible by the combination of these cycles. He simulated a large engine (40 MW) and a small engine ( 100kW), and evaluated their performance when coupled to a Lithium-Bromide (Li-Br) VARS, c onsidering both power and refrigeration as beneficial outputs. Auxiliary re frigeration was presented as a pe rcentage of nominal power. For the large engine, Boza estimated a thermal e fficiency of about 62% with 25% auxiliary refrigeration at 85F ambient conditions. This efficiency dropped only two points for the same operating point on a 103F day. Bo za points out that even higher e fficiencies are calculated if the auxiliary refrigeration is considered alongside of power in the efficiency definition. For the small engine Boza showed a thermal efficiency of about 43%, but with an additional 50% of auxiliary refrigeration, again on an 85F day. The thermal efficiency of this engine dropped three points for the 103F day. Boza also indicate d that ambient conditions directly affected only the VARS performance. This highlights the co mbined-cycle feature that refrigeration and

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24 efficiency can be interchangeable, and that auxili ary refrigeration can be shunted back into the HPRTE cycle to recover efficiency points if desired. The most recent modeling work on the combined HPRTE/VARS cycle was completed by Khan [8] in 2006. Khans model was intended to predict and optim ize the design-point performance of a medium-sized engine with conservative, but modern efficiencies and temperature limitations. Khans models added to Bozas work in that he accounted for the water condensation associated with the gas-path coo ling, and explored the implications of using different refrigerants. The figure of merit used was a linear combination of efficiency, auxiliary refrigeration, and water extraction. His model pr edicted an attainable thermal efficiency of 40.5% at an optimal low-pressure-s pool pressure ratio of two, and a turbine inlet temperature of around 2550F. Khan also showed that 1.5 pounds of water could be extracted from the cycle for every pound of propane fuel consumed. Khan s results are similar to Bozas small-engine results with the exception of hi s predicted external refrigera tion% of nominal power for a Li-Br cycle. This difference is probably relate d to Khans accounting for water condensation. Experimental Studies Several experimental research in itiatives have also been carri ed out at the University of Florida to further explore charac teristics of HPRTE performance and operation. One program in particular was funded by NASA in the late 1990s to evaluate a turbocharged Titan T62T32A gas turbine engine operating in a semi-closed cycle configuration. This program [9] proved many of the concepts upon which the HPRTE is f ounded. Though plagued by hardware problems unrelated to the focus of the research, HPRTE operations yielded power increases of up to 70% by turbocharging. Shortcomings in the fuel and control systems limited this test program from fully realizing the constant efficiency and SFC po tentials of the HPRTE. However, some simple scaling arguments provided within [9] show th at with more accommodating controls and fuel

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25 components, the efficiency and SFC would remain much more constant for various operating points. The anticipated, positive impact of semi-c losure on emissions also was validated within this test program. The HPRTE CO emissions were observed to be a factor of 25 less than the baseline values. This was thought to result from larger proportions of water in the working fluid due to recirculation. Spray coolers were also employed on the apparatus to aid in recirculationgas cooling, and also contribute d to a higher water concentrati on in the working fluid. The reduction of NOX emissions was also observed, but to a le sser extent than predicted by Muley in [6]. The potential for reduced NOX emissions was hindered by lowe r-than-expected recirculation flow rates and non-uniformities in the combustor te mperature field. Other experimental efforts, detailed in [10] and also funded by NASA, fo cused on the development of more suitable combustors for stable, semi-closed engine operations. Previous experimental efforts have, by necessi ty, incorporated some means of cooling to the recirculation gas path, as th is is essential for reliable and efficient engine operation. Most frequently, one or more air-to-water intercoolers were employed to meet this need. To date, based on the apparent absence of an y relevant literature, it seems as if no experiments have been performed on a semi-closed cy cle gas turbine combined with a VARS, so the present study represents the firs t of its kind.

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26 CHAPTER 3 TESTING HARDWARE An experimental apparatus was conceived for demonstrating the capabilities of the HPRTE/VARS combined cycle. This apparatu s was centered around a R over 1S-60 gas turbine engine and a custom-built VARS. In order to operate the combined cycle and successfully acquire useful data, a host of ot her components were necessary, such as facility cooling resources and instrumentation. This chapter describe s each of the elements in the HPRTE/VARS combined cycle, their roles, and thei r specifications, where available. Combined Cycle Components This section lists the various components re lating to the combined cycle, along with a detailed description for each. Engine Characteristics and Specifications The Rover 1S-60 is a single spool turboshaft engine that was designed and manufactured in the late 1950s and early 1960s. It was prim arily used for Auxiliary Power Units (APU) and fire-fighting water pumps before being used as an educational tool. The engine remained in production for about 20 years during which its safe ty, reliability, and long evity were well proven [11]. The engine utilizes a 19 blade, radial compre ssor and radial diffuser vanes with a design flow rate of about 1.33 lbs/s (0.603 kg/s), and pre ssure ratio of 2.8:1. To accommodate the dilute oxygen concentrations in the semi -closed cycle, the reverse-flow combustor was modified by welding stainless steel straps over some of the dilution holes, forcing more gasesand more oxygeninto the primary zone. Modifications were also made to intercept the flow from the compressor, redirecting it through the recuperato r before returning it to the combustor. The working fluid is then expanded acr oss a 30-blade axial turbine. Th e nominal rated output of this

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27 engine is 60 hp (45 kW) at a mechanically gov erned shaft speed of 46,000 rpm, with a thermal efficiency around 14%. The rotor speed is reduced in a gearbox on the front end of the engine to 3000 rpm, where a shaft transmits powe r to a water-brake dynamometer. Further modifications were also made to the engine casing, replacing it with thicker steel. This was necessary before pressurized operations could be safely attempted since the pressure within the casing was expected to increas e by as much as a factor of two. Recuperator The recuperator is a custom unit designed and fabricated by Elanco. It is a single-pass tube-and-shell heat exchanger with 672, 0.375inch (9.53 mm) stainless steel tubes extending 27 inches (69 cm) through the shell-side. The tube s are enclosed with 14inch (36 cm) diameter, 40 gauge stainless steel [12]. The hot-side inlet is attached to the turbine exhaust ducting with a Vband clamp, while the hot-side uses an 8-bolt flange to connect to downstream ducting. The cold-side ducting attaches with 12-bolt flanges. The recuperato r design effectiveness is about 0.51 [12]. Turbocharger The turbocharger used was a Garret GT 4294-731376-1, with a turbine scroll Area to Radius (AR) ratio of 1.44. This turbocharger was chosen primarily for its superb compressor efficiency, which would delivery higher pressure ra tios with relatively cooler turbine-side inlet temperatures. Additionally, the Garret 4294 wa s shown by an off-design model to match well with desired HPRTE operating points [13]. The tu rbocharger relies on an independent oil pump, dry sump, and oil cooler to provide its lubrication requirements. Heat Exchangers Three heat exchangers were made necessary by the recirculation of hot exhaust gases to the front end of the engine. All thr ee are arranged in series in the recirculation line, and each is

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28 named according to the qualitative state of th e recirculation gases entering it: the Hot Gas Cooler (HGC), Warm Gas Cooler (WGC ), and Cold Gas Cooler (CGC). Hot gas cooler (HGC) The HGC is a custom heat exchanger desi gned and fabricated by Energy Concepts in Annapolis Maryland for this a pplication. End-to-end, it is 36 inches (91 cm) long, and transitions from six inch Outsid e Diameter (OD) ducting, to a 12 inch (30 cm) OD shell, and back to the six inch ducting. No baffles are employed within the h eat exchanger, and all stainless steel construction is us ed. The tube-side heat transfer medium is a strong solution of ammonium and water. There is intended to be phase change within the HGC tube-side, since it serves as the vapor generator for the VARS. Warm gas cooler (WGC) The WGC was necessary if sub-ambient inlet temp eratures were to be possible. It serves the purpose of rejecting heat to ambient (though it uses a cooler-than-ambient medium) before recirculating gases are last cooled by the CGC. The WGC is an ai r-to-water heat exchanger that makes use of the facilities chilled water circuit. Chilled water makes two passes in the tube-side, while the recirculation gases tu rn through several baffles on one shell pass. Again, Elanco designed and manufactured the WGC using all stainless steel construction. The design effectiveness of this heat exchanger is 0.85 [12]. Cold gas cooler (CGC) The CGC was also constructed of stainless st eel by Energy Concepts, and functioned as the evaporator for the VARS. This heat exchanger is, in principle, capab le of chilling the High Pressure Compressor (HPC) inlet gases to sub-am bient temperatures. It is longer than the HGCabout 48 inches (1.2 m)with the same type s of transitions on each end, and also uses no baffles within. Ideally, the heat transfer medium is pure, liquid ammonia undergoing phase

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29 change with a few degrees of superheat upon ex iting the CGC. The gases on the shell side consist of recirculation gases mixed with fresh air from the Low Pressure Compressor (LPC). Dynamometer The dynamometer is contemporary with the R over 1S-60 and was sold as an integral package with the engine. It is a water-brake manufactured by Heenan & Froude (now Froude Hofmann) and is equipped with a mechanical tachometer. The spring balance with which the dynamometer was originally equipped was replaced with a load cell and reader. The original, mechanical tachometer is also supplemented with an optical tachometer directed at reflective tape on the Power Take-Off (PTO). Ducting The ducting used throughout the HPRTE rig wa s six-inch (15 cm) OD aluminized steel with a thickness of 0.070 inches (1.78 mm). It was manufacture d for use as exhaust pipe for large trucks, and was chosen fo r its availability, low cost, an d light weight. Pipe-to-pipe connections were made using 0.375 inch (9.53 cm ) thick, 12-bolt custom flanges which were welded to each pipe-end. Facility Resources There were several of the build ing assets that were both usef ul and essential for conducting engine runs. These dealt primarily with the issu e of removing heat from the engine room. The building chilled water circuit served to remove heat from the power and refrigeration cycles, while the ventilation systems available helped to maintain uniform temperatures within the EGDSL during engine runs. Chilled Water The building in which the EGDSL is situated utilizes a closed-circuit Process Chilled Water (PCW) system to function as a cooling me dium for several labs in the building. This

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30 system was tapped into and can be accessed from one of two headers in the EGDSL: one on the north wall, and one on the south wall. Each h eader is capable of de livering about 35 gpm (2.2 kg/s), or one header can deliver about 64 gpm (4.0 kg/s). All of the heat rejected from the recirculation gas path ultimately finds its way in to this cooling circuit. The VARS is cooled by the south header, and the WGC by the north header. The PCW circuit in turn rejects its heat to a stacked plate heat exchanger elsewhere in the bu ilding, which has a maximum heat transfer rate of almost 200 TR (703 kW). PC W temperatures are typically c ontrolled to be around 55 F (286 K), but by altering the PCW control scheme can be made as cool as 42 F (279 K). Fans and HVAC To impede the accumulation of harmful exhaust fumes within the testing facility, two fans have been employed to move fresh air through the EGDSL. An existin g air conditioning unit also exists in the EGDSL, the service of which is shared with a neighboring room. While this unit does little to cool the room during engine op erations, it does help to move more fresh air from elsewhere into the EGDSL. Vapor Absorption Refrigeration System (VARS) The VARS is a custom refrigeration unit de signed and built by Energy Concepts. The VARS is a unique, single-effect, ammonia-wate r absorption refrigeration system with a maximum evaporator heat load of 19 TR (67 kW), at a COP of 0.85. The VARS interacts with the HPRTE via the HGC and CGC described above, and rejects heat from the absorber and condenser through the south PCW header. Overview of Build Configurations To help ensure that all of the hypothesized at tributes of this combined cycle could be demonstrated and explored, several small, progre ssive steps were taken with the experimental apparatus. With some changes in the appa ratus different compone nts were chosen and

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31 implemented, and with other changes the same components were used, but in a different configuration. The configurati ons generally progressed such th at different capabilities were accentuated with different confi gurations, ultimately arriving at an apparatus capable of exploring all of the HPRTE/VARS combined cycle attributes. There were four different build configurations: Build 1, Build 2A, Build 3, and Build 4. Build 1 Build 1 was one of the first HPRTE testing platform s utilized in this series of tests. It was the configuration used for some of the previous research work discussed in Chapter 2. A block diagram of Build 1 is shown in Figure 3-1. The Build 1 configuration initia lly relied on a single, air-to-wat er heat exchanger to reject all of the heat from the recirculated ga s line. This is the HGC shown in Figure 3-1. Subsequently, a surrogate VARS evaporator was added to the recirculation flow path to investigate the effects of additional, low-temper ature recirculation cooling. Also, new water extraction hardware was designed and built to accommodate the newly addition condensation location. The evaporator was part of a small, vapor compression refrigeration system capable of yielding about one ton (3.5 kW) of cooling. It was installed within a 0.250-inch (6.35 mm) thick steel box that previously served to house a filter. A photograph of this installation is shown in Figure 3-2. The improvised evaporator yielded limited su ccess in furthering the water extraction for Build 1. However, an important point was made clearer. Significant water extraction could still be attained without cooling all of the recirculated gas mixture down to its saturation temperature. The cold tubes accomplished the local cooling of gases near to the tube surfaces, and successfully condensed apprecia ble quantities of water.

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32 Overall, Build 1 tests were successful in pr oving the concept of wa ter extraction from the HPRTE. However, no efforts were made to pressurize this engi ne configuration. Build 2A It was believed that the sec ond phase of this research w ould permit pressurized engine operations, and greater gas-path cooling. Before moving prematur ely to a ducting configuration incorporating a VARS, more simulated VARS coo ling was used instead. This decision was influenced by the availability of two air-to-water heat exchangers in the EGDSL. These could be implemented practically, allowing more to be lear ned about recirculation cooling in a reasonable amount of time. This successive ducting configur ation was referred to as Build 2A. A block diagram of the Build 2A configuration is provided in Figure 3-3. Build 2A utilized two, tube-andshell heat exchangers to re ject gas-path heat to the building PCW circuit. A summative 40-50 TR ( 140-175 kW) was consistently removed from the recirculation line with these two coolers. This helped to demonstrate some of the concepts related to VARS integration since the means by which the recirculation gases are cooled are arbitrary. These two heat exchangers simulate d the VARS vapor generator and evaporator to some extent, though the cooling medium (PCW) was different. In addition to yielding exceptional water extr action results, the Build 2A configuration provided some valuable information about pressu rized operations. It was observed that some means of throttling is necessary between the LP T inlet and LPC exit in order to pressurize. Otherwise, the exhaust gases may flow the wrong way through the LPC, out to ambient, since, depending on the ducting configur ation, this can be the less resi stive path. Having this fact reinforced experimentally was f ound useful when considering th e next build configuration.

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33 Build 3 The next build configuration was called Build 3. In this configurat ion, one of the air-towater heat exchangers was replaced with the vapo r generator of the VARS. Also, one more heat exchanger, the CGC, was added to the recircula tion line. The CGC is the evaporator in the VARS and provides the opportunity to condense more water than the previous build with a lower temperature heat transfer medium. Furthermore, Build 3 was the first configuration to utilize the new, modified Rover 1S-60 engine. The new e ngine was equipped with a much thicker casing made necessary by higher pressure operations. It was feared that the Rover 1S-60, as originally manufactured, would have ruptured with the greater HPC exit pressures within. Lastly, the Build 3 configuration benefited from a new Garret 4294 turbocharger. This tu rbocharger was chosen because it was more capable of higher pressure ra tios, given the same, relatively cool recuperator exit temperature. With the previous turbocharg er, models showed that pressure ratios of 2:1 were unattainable with the relatively low LPT inlet temperature. Plans were originally in place to duct some of the hot, High Pressure Turbine (HPT) exit gases around the recuperator, but the replacement of the old turboc harger with newer technology wa s more practical. A block diagram of the Build 3 configuration is shown in Figure 3-4. One characteristic of the Build 3 configuration was that the LPC exit gases mixed with the recirculation gases immediately downstream of th e CGC. This was bene ficial in that the pressure drop between the LPC exit and HPC inlet was minimal. However, the drawback to this mixing location was that the LPC exit flow was cooled only by mi xing with the cooler recirculation gases. At higher LPC pressure ratios, and LPC exit temperatures, the mixed temperature of the HPC inlet gases would be a pproaching the limit imposed by the manufacturer. In consequence, a move to the final configuration was made.

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34 Build 4 The final configuration, Build 4, involved onl y a small change in the gas path. The abovementioned LPC exit mixing location was move d to an elbow directly upstream of the CGC. By so doing, the mixed recirculated e xhaust gases and fresh LPC air would all pass through the CGC together to be cooled. Initially, this may seem to be no different, in a thermodynamic sense, as the same quantity of heat is removed from the gas path whether mixing is done before or after the CGC. However, by mixing before the CGC, the flow rate, and Reynolds number passing through it incr eased from 60% to 100%, depending on the recirculation ratio and LPC flow rate. Moreover, the CGC inlet temperature will generally be increased by the mixing, since the LPC exit temper atures were usually higher than the WGC exit temperatures. These two effects, higher Reynol ds number and higher inlet temperature, were expected to increase the effectiveness of the CG C. Thus, for the same conditions, the HPC inlet temperatures were expected to be lower for Build 4 than Build 3. A block diagram of Build 4 is shown in Figure 3-5. Data Measurement Testing of the HPRTE/VARS combined cycle necessitated the use of various types of instrumentation. This section describes in deta il all of the hardware and software used to measure and acquire experimental measurements. Data Acquisition Hardware The Data AcQuisition (DAQ) hardware is he re defined as the equipment between all transducer wire terminations a nd the data output. Hence, ther e are two component s that provide this throughput of information: a National Inst ruments chassis, and a personal computer. The purpose of the DAQ hardware is to receive the ra w, analog signals from various transducers and, after filtering and amplif ication, convert each of them into a digital signal. From this point a

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35 1200 MHz Personal Computer (PC) operating with 256 Mb of RAM received the digital stream and readied it for the DAQ software. The DAQ hardware used for these experiments c onsisted of one chassis, four modules and four terminal blocks. The chassis is a mode l SCXI 1000, and provided a low-noise environment in which signal conditioning can be performed by the modules. A 50-pin cable conveyed the data to a PCMCIA card installed within the PC. Two SCXI 1102 modules were employed for th e acquisition of thermocouple signalsone for the HPRTE temperatures, and one for th e VARS temperatures. The SCXI 1102 modules were equipped with 32 differential, analog ch annel inputs, not including one Cold-JunctionCompensation (CJC) channel. Additionally, each channel was amplified with a gain of either one or 100 at the discretion of the user. Each channel in the 1102 modul es was also equipped with a three-pole lowpass filter re jecting 60Hz noise. SCXI 1300 terminal blocks were used with each of the SCXI 1102 modules for the thermocouple wire terminations. One SCXI 1100 module was utilized to condition all of the pressure tr ansducer, load cell, and tachometer reader signals. The SCXI 1100 al so had 32 differential channels and an onboard programmable gain instrument am plifier. All channels were also equipped with a jumperselectable (four Hz or ten kHz cutoff), single-po le filter. Wiring for this module was terminated in a SCXI 1303 terminal block. One SCXI 1126 module was necessary for the acquisition of all inst rumentation having a frequency output. It was equippe d with eight isolated channels with filtering, and a software programmable frequency-to-voltage conversion circ uit. This module used a SCXI 1327 terminal block for its wire terminations. Within this te rminal block was a selectable 1:1 or 100:1 voltage attenuation switch for each channel.

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36 Data Acquisition Software The software used for processing the data si gnals was LabVIEW 7.1. It is also a National Instruments product, and therefore interfaces ve ry well with the DAQ hardware. LabVIEW is a graphical programming environment that allows instrumentation signals to be monitored, manipulated or scaled, and appended to a data file all in real-time. This type of interface was essential to monitoring key parameters on both the power and refrigeration sides of the combined cycle. Instrumentation In general, five types of measurements were made in testing the HPRTE/VARS: temperature, pressure, flow ra te, rotational speed, and load. Thermocouples were exclusively used for temperature measurements. For pre ssure measurements, diaphragm-type pressure transducers were implemented to provide the DAQ with an analog signal. In addition to these, redundant manometers and mechanical pressure ga uges were used for some of the more vital pressure measurements. Where measured direct ly, liquid flow rates were acquired by either turbine or paddle-wheel, selfpowered flowmeters. The only rotational speed measurement, dynamometer speed, was made with an optical tachometer, and load measurements were performed with load cells. Thermocouples Three different types of thermocouples were used for temperature measurements: J-type, K-type, and T-type. In general, the K-type, J-type, and T-type thermocouples measured hot gas, warm gas, and cool liquid temperatures, respec tively. Exceptions to this are on the VARS, where T-type thermocouples were used for a ll temperature measurem ents. Most of the thermocouples were 0.250 inches (6.35 mm) in diameter, and extended to approximately the centerline of the flow path.

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37 Pressure transducers The pressure transducers used were all Omeg a PX138 series sensors. These transducers employed an eight VDC power source with a on e VDC to six VDC output, and were equipped with temperature compensation up to 122 F (323 K). There were four types of differential PX138 sensors used, classified by pressure range : 0-100 psi (689 kPa), 0-30 psi (207 kPa), 0-5 psi (34 kPa), and 0-1 psi (7 kPa). All of thes e sensors were installed near the ceiling of the EGDSL, where shielded cables then carried thei r signals to the SCXI 1303 terminal block. Turbine and paddle-wheel flowmeters The fuel flow rate measurements were made using a Hoffer Flow Controls, low-flow liquid turbine flow meter, model numbe r MF1/2X70B. It was situated immediately downstream of the fuel tank and generated a freque ncy output in proportion to the volumetric flow rate. For a constant viscosity, this flowmeter establishes a li near response, which is transmitted directly into the SCXI 1327 terminal block. After correcting for an initial offset, accuracies within around 1% are typical from this instrument. Two paddle-wheel flow meters were used to measure the volumetric flow rate of PCW one at each header in the EGDSL. These units were ma nufactured by Omega Engineering, model number FP-5300. These instruments came with CPVC pipe fittings that would ensure the proper number of pipe diameters were in place, both upstream and down stream of the flowmeters. These flowmeters also generated a linear frequency output over volumetric flow rate. Optical tachometer Dynamometer speed measurements were accomplished with a Monarch ROS-5W optical sensor and ACT-3 reader. The sensor was pla ced about 0.5 inches (1.3 cm) from the PTO, the circumference of which was covered with electri cal tape. A single strip of reflective tape was

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38 placed over the electrical tape, triggering the optical sensor once per dynamometer revolution. The reader displayed the dynamometer speed in the control room, but al so provided a 0-5 VDC output, from which the DAQ hardware received its speed signal. Load cells Two load cells were used during engine tests. The more important of the two was used to counter the dynamometer torque. This instrument was necessary for measuring the engine load, and in turn the net power output. The output from this load cell was interpreted and displayed by an Omega Engineering DP-145 multi-purpose reader Unfortunately, the specifications for this particular load cell could not be located. The second load cell was employed to measure the quantity of water being extracted from the HPRTE in real-time. A reservoir was suspended from this instrument, which was cantilevered from a small support structure. Se veral hoses conveyed the condensate from within the ducting into the reservoir over which they were suspended, also from the support structure. Thus, the real-time weight of the extracted wate r could be straightforw ardly differentiated to obtain a flow rate. This particular load cell wa s an Omega Engineering platform load cell with a 72 lb (33 kg)capacity. Gas analysis A Cosa 1600 IR analyzer was used to evaluate the composition of exhaust gas samples taken from the exhaust before leaving the EGDS L. It is a portable, five-gas exhaust gas analyzer. It quantifies carbon monoxide (CO), carbon dioxide (CO2), and unburned hydrocarbon (UBHC) concentrations using non-dispersive infrared technol ogy. Diatomic oxygen (O2) and nitric oxide (NO) concentrati ons can also be measured using electrochemical sensors.

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39 Mechanical instruments and analog readers The HPRTE/VARS experiments did not rely exclusively on the DAQ system for all values, but also utilized redundant sets of manua lly recorded data. This practice was adopted primarily to insure against the complete loss of data in the ev ent of digital data corruption. Another benefit from practicing manual data reco rding stems from the fact that much of these data, particularly the pressures and differentia l pressures, were measured with different mechanical instruments. In these cases, the analog and digital sets of values could be compared to mutually ensure the reliability of both. However, this wasnt the case for the temperature measurements since the signal from each thermoc ouple bifurcated to an analog reader and the DAQ. Consequently, both values were always the same. Each type of thermocouple had its own multip lexer and Omega DP460 reader. Thus, three multiplexer/reader pairs were implemented to di splay all of the HPRTE and VARS temperatures. Specifics on each of the temperature measuremen t locations and their respective thermocouples are provided in Table 3-1. The state points correspond with those shown in Figure 1-1. All of the HPRTE pressure taps branched off to a redundant gauge of some sort. The more important of these pressures, deem ed vital to engine health duri ng experiments, were routed to their own individual gauge. This improved the visi bility of important pressures. Other, less critical pressure hoses were each connected to one of two headers through its own valve. Each header was connected to a diffe rent Bourdon tube pressure gauge The pressure panel operator could choose to view a specific pressure by si mply opening the valve for that hose. After viewing and recording a pressure, the panel opera tor would then close the valve, and vent the header. The header would then be ready to recei ve a new pressure. Details on each of the state point gauges are shown in Table 3-2.

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40 All of the differential pressures were treate d similarly. Each diffe rential pressure line going to a transducer also teed off to some sort of manometer. Manometers of various sizes and shapes were used, employi ng sundry fluids. Table 3-2 also presents the specifications for each of the manometers used, alongside the specific diffe rential pressures for which they were used. Lastly, the fuel flow rate utilized an addi tional rotameter, inline with the turbine flow meter. This device was always viewed helped validate the weak, selfpowered signal of the turbine flowmeter. The rotameter outputs a mass flow rate, and is calibrated for a specific gravity of 0.835. Table 3-1. Temperature locations and instrumentation. State Point Location Thermocouple TypeInstrument Range Expected Range of Values 0 Ambient J -40-1380F (251-1022 K) 40-100F (278-311 K) 1 LPC inlet J -40-1380F (251-1022 K) 40-110F (278-316 K) 2A LPC exit J -40-1380F (251-1022 K) 100-250F (311-394 K) 2.9 CGC exit J -40-1380F (251-1022 K) 40-120F (278-322 K) 3 HPC inlet J -40-1380F (251-1022 K) 40-120F (278-322 K) 4 HPC exit J -40-1380F (251-1022 K) 300-380F (422-466 K) 5 HPR inlet J -40-1380F (251-1022 K) 350-420F (450-489 K) 6 HPR exit J -40-1380F (251-1022 K) 580-620F (578-605 K) 7 COMB exit K -320-2200F (77-1478 K) 1300-1600F (978-1144 K) 9 HPT exit K -320-2200F (77-1478 K) 900-1200F (755-922 K) 9.1 LPR exit K -320-2200F (77-1478 K) 780-1100F (689-866 K) 9.2 LPT inlet K -320-2200F (77-1478 K) 780-1100F (689-866 K) 9.3 WGC inlet J -40-1380F (251-1022 K) 500-600F (533-589 K)

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41 State Point Location Thermocouple TypeInstrument Range Expected Range of Values 9.4 WGC exit J -40-1380F (251-1022 K) 110-185F (316-358 K) 10 LPT exit K -320-2200F (77-1478 K) 800-900F (700-755 K) 1w PCW supply T -320-660F (77-622 K) 42-75F (279-297 K) 2w PCW return T -320-660F (77-622 K) 42-85F (279-303 K) 3w PCW supply T -320-660F (77-622 K) 42-75F (279-297 K) 4w PCW return T -320-660F (77-622 K) 42-85F (279-303 K) Table 3-2. Pressure locati ons and instrumentation. State Point Location Type of Instrument Instrument Range Expected Range of Values 0 Ambient Internet Resource N/A 736-788 mm-Hg 0-1 LPC inlet dP Angle Manometer SG 0.827 0-2 in-H2O (0-4 mm-Hg) 0-2 in-H2O (0-4 mm-Hg) 0-2B MAI dP Angle Manometer SG 1.91 0-7 in-H2O (0-13 mm-Hg) 0-7 in-H2O (0-13 mm-Hg) 2A LPC exit Bourdon Tube Gauge 0-60 psig (0-414 kPa) 0-15 psig (0-103 kPa) 2.8-2.9 CGC dP Manometer SG 1.0 0-60 in-H2O (0-112 mm-Hg) 0-48 in-H2O (0-90 mm-Hg) 3 HPC inlet Bourdon Tube Gauge 0-60 psig (0-414 kPa) -0.5-15 psig (-3.4-103 kPa) 4 HPC exit Bourdon Tube Gauge 0-100 psig (0-689 kPa) 0-65 psig (0-448 kPa) 5 HPR inlet Bourdon Tube Gauge 0-200 psig (0-1379 kPa) 0-65 psig (0-448 kPa) 6-7 COMB dP Manometer SG 1.75 0-60 in-oil (0-197 mm-Hg) 0-50 in-oil (0-164 mm-Hg) 8 HPT inlet Bourdon Tube Gauge 0-200 psig (0-1379 kPa) 0-60 psig (0-414 kPa) 9 HPT exit Bourdon Tube Gauge 0-60 psig (0-414 kPa) 0-20 psig (0-138 kPa) 9.1 LPR exit Bourdon Tube Gauge 0-200 psig (0-1379 kPa) 0-20 psig (0-138 kPa) 9.3 WGC inlet Bourdon Tube Gauge 0-200 psig (0-1379 kPa) 0-20 psig (0-138 kPa)

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42 State Point Location Type of Instrument Instrument Range Expected Range of Values 9.5 RCV dP U-Tube Manometer SG 0.827 0-70 in-oil (0-109 mm-Hg) 0-65 in-oil (0-101 mm-Hg) 10 LPT exit Bourdon Tube Gauge 0-200 psig (0-1379 kPa) 0-20psig (0-138 kPa) Figure 3-1. Block diagram of Build 1 configuration.

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43 Figure 3-2. Photograph of Build 1 im provised evaporator installation. Figure 3-3. Block diagram of Build 2A configuration.

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44 Figure 3-4. Block diagram of Build 3 configuration. Figure 3-5. Block diagram of Build 4 configuration.

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45 CHAPTER 4 TEST PROCEDURES Test Plan Before conducting an engine test, a test plan needed to be prep ared so that time and fuel were not wasted making decisions during the run. Fist, the test plan identified the operating points to be evaluated. Then a plan was c onceived to achieve each operating point since transitions into the pressurized regime were onl y successful after taking the proper steps in the correct order. Such a plan usually consisted of the gradual closing of the boost control valve (VEXH in Figure 1-1), while carefully maintaining a reason able recirculation ratio and fuel-to-air ratio using the recirculation valve (VREC in Figure 1-1). Contingency plans were also needed in cas e a particular operati ng point could not be reached. It was often the case, for example, that a desired pressure rati o could not be reached using the test plan. Instead of shutting down th e engine and allowing the days efforts to go to waste, it was deemed prudent to have a sec ondary objective for the same engine run. An example of some operating procedures is provided in Appendix A. Engine/Component Preparation Prior to starting the HPRTE and VARS, it was e ssential that several, pre-planned steps be taken to ensure personnel safet y, engine longevity, and data in tegrity. A thorough checklist was created for each engine run, and rigorously reviewed before that run. Quantities such as oil level, battery charge, PCW flow rate, etc. were confirmed as acceptable. Following these preparations, a regular startup procedure was em ployed in an attempt to further remove any question regarding the consistency and integrity of all engine runs An example of these setup procedures is provided in Appendix B.

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46 Instrument Preparation Additional actions were taken before e ngine runs specifically to prepare the instrumentation, DAQ hardware, and DAQ software Some checks were made before every engine run, while others were less frequent. Ho wever, all were deemed necessary for acquiring consistent, accurate, and controlled data from the HPRTE/VARS combined cycle. Transducer Offsets Since the DAQ hardware was equipped with CJC capabilities, there was little maintenance necessary for the thermocouple measurements. However, infrequent and random tests were employed to make certain that thermocouples were maintaining their accuracy and precision. Thermocouples were immersed in a small pool of boiling water away from the heated surface to ascertain whether or not the pr oper boiling temperature was being output. Thermocouples rarely deviated more than 1 F (0.6 K) or so from the expected temperature. If one particular thermocouple was somewhat off at this temperatur e, then an offset wa s added to that signal within the DAQ software. Such a simple assessment may be flawed in that it is only performed at one temperature in the thermocouple range. It was assumed that th e thermocouple linearity was retained, since calibrations at higher temperatures were not practical. The pressure transducers also required that meas ures be taken to ensure consistent outputs. Periodically, the transducer calibrations were checke d to verify that at le ast their pressure-overvoltage slope had remained the same. Any offs ets in their outputs were addressed at the conclusion of engine runs, but before analysis of the data. The linear calibration line frequently shifted up or down by a small amount. To correct for this several minutes of lead-in da ta were recorded before each e ngine startup. During the lead-in interval, all of the pressu re transducers should have indicated a gauge pressure of zero. The leadin data was averaged, and then s ubtracted from all subsequent data points for that engine run.

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47 The intended result was that each pressure measur ed as zero psig (0 kPa) before startup and following shutdown. The pressure transducers were powered with a variable output DC power supply. All transducer calibrations were perf ormed with an excitation voltage of 8.11 VDC. Any drifts in this output were compensated for within the DAQ software. Before each raw voltage output was translated to a pressure, the signal was scaled according to the real-time excitation voltage (also read into the DAQ) relative to the calibration volta ge. This effectively co mpensated and drift in the excitation voltage, since the transducer output voltages scale one-to-one with the excitation voltage. Ambient Conditions The performance of any gas turbine engi ne depends greatly on the thermodynamic conditions of its environment. As such, the am bient temperature, pressure and humidity were measured or calculated. The ambient temper ature was obtained from one of two J-type thermocouples. The first of these was mounted in front of the Main Air Inlet (MAI) valve (VMAI in Figure 1-1). However, for most of the Build 4 engi ne runs this valve wa s closed and none of the HPRTE working fluid moved over that th ermocouple. Thus, the second of these thermocouples, mounted in front of the LPC inle t bellmouth, was most often used to represent the ambient temperature. The two thermocoupl es usually measured different temperatures because of their locations. This resulted fr om the way in which air was drawn through the EGDSL. Fresh air from outdoors moved over the MAI thermocouple first, and then flowed over and around some of the hot ducting, heat exchange rs, and engine surfaces before reaching the LPC inlet. Accordingly, the LPC inlet flow wa s a few degrees warmer than the MAI flow. No barometers were used to directly measur e the atmospheric pressure in the EGDSL. This value was obtained from the University of Florida Department of Physics Weather Station,

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48 which publishes this data online. These valu es are updated every five minutes, and were recorded at the beginning and e nd of an engine run. The Weathe r Station website also stores 24 hour trends in case knowledge of barometric pres sure variations during the run is desired. The wet and dry bulb temperatur es were measured directly within the EGDSL using a psychrometer. This measurement was repeated at the beginning of each trial during an engine run. At the conclusion of each run these temperat ures were used with a psychrometric chart to obtain the humidity ratio and rela tive humidity for each trial.

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49 CHAPTER 5 DATA REDUCTION SCHEME During the four months of engine testing, m illions of data were acquired, along with the challenge of analyzing them. For the period of each engine run, ther e were 63 values being measured and recorded, once every 1.5 to two sec onds. Each of these 63 pressures, temperatures and flow rates needed to averaged, corrected wh ere necessary, and then pr ocessed to arrive at performance parameters such as power, efficien cy, heat load, etc. Thus, an algorithm was developed within the Matlab environment to carry out all of the statis tical operations, and the reduction and organization of data. The met hods employed to reduce the HPRTE data are detailed below. However, those measurements obtained from the VARS instruments were used exclusively for VARS operations, and were too few to suffice for a performance analysis. Consequently, only the power cy cle data were analyzed. As with any engineering analyses, some a ssumptions and approximations were necessary in order to have a complete set of equations. Since many of the desired performance parameters cannot be measured directly, each thermodynamic pr ocess and substance needed to be classified so that it would receive the proper treatment. Additionally, the geometric, fluid, and heat transfer properties of various components were of ten too complicated for the application of some analyses, further compelling the use of good engineering assumptions. Gas Properties Given the primary role played by the working fluid in a power cycle, it is necessary to describe how these fluids are analytically tr eated. The HPRTE presents a unique flow path wherein many various thermodynamic processes take place: mixing of ga ses, power transfer to/from gases, heat transfer to/from gases, in addition to the condensation and subsequent

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50 extraction of water. This section will descri be the assumptions, approximations, and equations specific to the HPRTE gases. Enthalpy and Specific Heat The working fluid within the power cycle wa s approximated as a thermally perfect gas, composed of N2, O2, CO2, water vapor (H2O), and CO, and as behaving in accordance with the perfect gas law shown in Equation 5.1. T R PMIX (5.1) Here the mass-specific gas constant, RMIX, is computed with Equations 5.2, using knowledge (or assumed knowledge) of the chemical co mposition at a giving state point: MIX MIXM R R (5.2a) i i i MIXM y M (5.2b) where yi represents the mole fraction of the ith species being considered, Mi represents the molecular weight of the ith species, and MMIX represents the molecular weight of the mixture. The variable, R is the universal gas constant. For none of the calculations was the working fluid approximated as calorically perfect. Instead, a third-degree polynomial was used to ob tain the constant-pressu re specific heat (cp) of a given gas at a given temperature. The system used for obtaining cp for one of the pure gases comprising the mixture is shown in Equations 5.3 [14]. 3 21 c b a c cpo p (5.3a) 1 oT T (5.3b)

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51 In Equations 5.3, the coefficients, a, b, and c, are part icular to the pure gas being considered, and cpo is the constant pressure specific heat of the pure gas at the reference temperature (To) of 300K. In Equation 5.3b, T is the temperature of the pure gas, which is assumed to be the same as the mixture temperature. Following th e computation of each constituent cp, that of the gas mixture could be calculated using Equation 5.4. i i i P i MIX MIX PM c y M c, ,1 (5.4) The above describes only how, when necessary, the cp was computed for a gas mixture at a single, given temperature. However, the mo re pertinent and more frequently occurring computations were the specific enthalpy changes ( h) within the various components. The h calculations were carried out di rectly by combining Equations 5.3 and integrating definitely over the measured temperature difference, per Equation 5.5. Tout Tin o o o o o po Tout Tin pT T cT T T bT T T T a T c dT c h 4 3 21 4 1 3 2 (5.5) Here again, the h on the left-hand side of Equations 5.5 is for a single species within the gas mixture, and needs to be computed for each of the five constituents. Finally, the h of the gas mixture could be calculated using Equation 5.6. i i i i MIX MIXM h y M h 1 (5.6) It was often necessary to make use of an average cP across certain of the thermodynamic processes. For example, some of the turbomach inery with large temperature differences needed an average specific heat ratio to use in efficiency calculations. For cases in which a temperature difference was measured, and h calculated, Equations 5.7 were used to obtain an average constant pressure specific heat (cP,AVG), and specific heat ratio ( AVG).

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52 T h cMIX AVG P (5.7a) AVG P MIX AVGc R,1 1 (5.7b) Gas Composition Specialized instrumentation was employed to measure on the composition of gases throughout the HPRTE flow path. It was impractical, however, to directly measure gas samples from more than one location, because of th e time needed to acquire just one sample. Consequently, assumptions were made allowi ng the gas composition to be followed through various, HPRTE processes effecting changes in th e gas mole fractions. One substance of chief importance in this study was the content of wate r vapor contained within the working fluid. Since the HPRTE water extraction capability so strongly motivated this experimental endeavor, both the quantity of water extracted and the quantity not extracted were considered as key performance parameters. A problem presented by the gas analyzer was that it failed to include a direct quantification of water vapor in the samples it measured. The analyzer only assessed the concentration of the following species: O2, CO2, CO, NO, and a methane equivale nt of UBHCs. A persisting question was whether water vapor is considered by the machine when these concentrations are output, or if they are dry concen trations. Unfortunately, this question was never satisfactorily answered by the gas analyzer OEM, and some assumptions were necessary to complete the analysis. It was assumed that the gas anal yzer provided dry concen trations, according to Equation 5.8, UBHC NO CO CO O N TOTn n n n n n n 2 2 2 (5.8) where nTOT represents the total number of moles bei ng considered in a given sample. However

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53 for this analysis, as it pertai ned to thermodynamic processes, the quantities of NO and UBHCs were ignored, since they were several orders of magnitude smaller than the other constituents. The CO mole fraction was also ve ry smallless than 1%but was retained for later service in determining combustion efficien cy. Given the absence of N2 concentrations from the gas analyzer output, Equation 5.9 was used to calculate the dry N2 mole fraction, completing knowledge of the dry exhaust gas composition. DRY CO DRY CO DRY O DRY Ny y y y, 2 2 21 (5.9) Attention was then turned to the estimation of water content in each of the gas analyzer samples, beginning with the assumption that the gas mixture was saturated upon reaching the analyzer. Affirming this assumption was the cons istent presence of condens ate droplets inside a transparent trap on the analyzer. Furthermore, th is trap was consistently at or slightly above room temperature. Combined with atmospheric pr essure data, the partial pressure of water vapor within the exhaust gases, PH2O, was then estimated using saturated water properties from [15, pp. 924-925], providing the water mole fraction. ATM O H O HP P y2 2 (5.10) Once again assuming the exhaust gases to behave as an ideal gas mixture, the summative partial pressures of the dry constituents, PTOT,DRY, was determined with Equation 5.11. O H ATM DRY TOTP P P2 (5.11) Lastly, each of the dry mole fractions, yi, were corrected to include th e water vapor, as shown in Equation 5.12,

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54 ATM DRY TOT DRY TOT DRY i iP P P P y, (5.12) where Pi,DRY represents the partial pressure of the ith dry exhaust gas constituent. Basic Equations and Assumptions The overall data reduction scheme was compos ed of a series of analyses performed on each of the HPRTE components. In additi on to the more obvious components such as compressors and heat exchangers, various se ctions of ducting are also analyzed as components, since many of these have some ki nd of thermodynamic process occurring within them. This section will provide a descrip tion of how each of the HPRTE components was analyzed, including equations a nd assumptions utilized. Main Air Inlet (MAI) The MAI admitted air to the HPRTE during star tup. It is an elliptical nozzle (or bellmouth) with a 15.3 square-inch (98.7 cm2) throat area (AMAI). A pressure tap at the nozzle throat was used to measure the differential pressure, PMAI, to atmosphere, and a J-type thermocouple measured the temperature of the inlet air, TMAI. Atmospheric pressure and relative humidity were also measured to ensu re use of the proper gas constant, RMAI. While the MAI was used only as a metering device for some of the HPRTE airflow, it was closed and admitted no air for the majority of the Build 4 trials. Since the maximum Mach numbers at the throat were around 0.17 the air mass flow rate was arrived at using incompressible flow theory, and is shown below in Equation (5.8). MAI MAI MAI ATM MAI MAIT R P P A m (5.8)

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55 Low Pressure Compressor (LPC) The LPC also employed a bellmouth at its inlet, as well as a pressure tap connected to instrumentation and a thermocoupl e at the inlet. The cross-se ctional area of the LPC bellmouth throat is 17.7 square-inches (114 cm2). The method of calculating the mass flow rate is shown in Equation 5.9, and again derives from incompressible flow theory. LPCI LPCI LPCI ATM LPCI LPCT R P P A m (5.9) The LPC power (LPCW ) was calculated using Equation 5.10: LPC LPC LPCh m W (5.10) where hLPC was computed via the methods outlined in the Gas Properties subsection above. The efficiency of the LPC was calculated from the definition of isentropic efficiency: 1 11 LPCI LPCX LPC LPCT T PRLPC LPC (5.11) where PRLPC is the pressure ratio across the LPC, TLPCX represents the LPC exit absolute temperature, and TLPCI represents the LPC inlet absolu te temperature. The variable, LPC, denotes the average specific heat ratio across the LPC, per the Gas Properties section above. Recirculation Venturi (RCV) The RCV was used in the recirculation flow pa th exclusively as a metering device for the recirculating gas flow. It was situated after th e WGC in a straight run of 6-inch (15 cm) ducting where the upstream pressure (PRCVI) and temperature (TRCVI) were measured, as well as the differential pressure ( PRCV) from upstream of the venturi to the throat. The throat area was 6.605 square-inches (42.61 cm2). Like the two inlet bellmouths on the HPRTE, principles from incompressible flow theory were used to compute th e velocity at the throat of the venturi, and in

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56 turn, the mass flow rate. The e quation used is shown in Equation 5.12, 1 2 1 2 21 A A T R P P P A mRCVI RCV RCV ATM RCVI RCV (5.12) where A1 and A2 are the inlet and throat areas, respectively. A primary operating parameter, the recirculati on ratio, was then defined as the ratio of recirculating gases to fresh air, as shown in Equation 5.13. MAI LPC RCVm m m R (5.13) Hot Gas Cooler (HGC) The hot-side of the HGC was bounded by tw o, K-type thermocouples and had approximately the same gas flow rate as that measured by the RCV. The difference in gas flow rate is small, but worth noting in Equation 5.14 as follows: WGC O H RCV HGCm m m, 2 (5.14) where WGC O Hm, 2 accounts for the water extraction from the flow path before reaching the recirculation venturi, but after the HGC. The heat load on the HGC hot side is computed as shown below. HGC HGC HGCh m Q (5.15) No attempt was made to compute the effectiveness of this heat exchanger, as insufficient data existed on the cold side of the HGCnamely the co ld side flow rate an d fluid properties. Warm Gas Cooler (WGC) Both the hot and cold sides of the WGC were instrumented with thermocouples. The coldside utilized a paddlewheel flow meter, and the hot-side flow rate was known to be the same as that of the HGC. Accordingly, both the WGC gas-side heat rate (WGCGQ ) and water-side heat

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57 rate (WGCWQ ) could be calculated as follows: WGC fg WGC O H WGC HGC WGCGh m h m Q, 2 (5.16a) w W p W W WGCWT c V Q (5.16b) In Equation 5.16b, w is an assumed water density of 62.4 lbm/ft3 (1000 kg/m3), WV is the measured volumetric water flow rate in ft3/s, cP,W is an assumed constant pressure specific heat for water of 0.998 Btu/lbm-R (4187 J/kg-K), and TW is the measured temperature rise of the cooling water. In Equation 5.16a, hfg,WGC, is the heat of vaporization. The heat of vaporization was found using data adapted from [15, pp. 924-925] after computing the water partial pressure, as shown below: ATM WGCI O H O HP P y P 2 2 (5.17) where yH2O is the mole fraction of water vapor in the gas stream, and PWGCI is the inlet pressure of the WGC in units of psig. Again in Equation 5.16a, WGC O Hm, 2 denotes the quantity of water condensed in and extracted from the WGC. Importan t to note at this point is that only the total water extraction rate was directly measuredt he summative water extracted from both the CGC and the WGC. Because there were no flow meas urements for the individual streams, a crude, qualitative fraction was applied to the measured, total extraction flow rate, which estimated the quantity of condensate issuing from each extr action point. For example, this fraction was typically 60/40, for the WGC/CGC extraction rates. It was most often the case that the two heat rates shown in Equations 5.16 were not equivalent. The gas-side heat rate always tended to be higher than the water-side. This was attributed to heat losses to ambient, which was straightforwardly found by adding the two heat rates.

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58 Efforts were made to calculate the effectiven ess of the WGC. The water condensation was one of the chief objectives motivating this experi mental program. Thus, the proficiency of the WGC to condense water was regarded as one of the HPRTE performance parameters, and was accounted for in the effectivene ss calculations, as detailed be low. The effectiveness was calculated as shown, WGC WGC WGCQ Q emax, (5.18) where WGCQmax, is the maximum possible gas-side heat rate, as defined by Equation 5.19. WGC fg O H SEN O H WGC NC SEN NC WGC WGCh h m h m Q, 2 max, 2 max, max, max, (5.19) In Equation 5.19, NC WGCm, is the non-condensable gas flow rate, hmax,SEN,NC is the change in enthalpy if the non-condensable gases are cooled to the water-side in let temperature, and O H WGCm2 max, is the estimated, maximum possible water ex traction rate if the mixture is cooled to the cold-side inlet temperature. The variable hmax,SEN,H2O represents the sensible change in enthalpy from cooling the water vapor to its initia l dew-point. Again, this treatment effectively considers the flow as two sepa rate streamsone condensable stream and one non-condensable streamand rests on the assumption of an ideal ga s mixture, even for the water near its vapor dome. Furthermore, it is understood that th e interaction of the wa ter vapor and the noncondensable stream will produce effects not capture d by this analysis. Specifically, the water vapor will experience further se nsible enthalpy changes below th e initial dew-point, since the dew-point temperature will gradually decrease as more water is taken from the mixture. Cold Gas Cooler (CGC) A mixing junction was situated directly upstream of the CGC where fresh air from the LPC joined with recirculating exhaust gases. Mixing junction calculations were necessary to find the

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59 mixed, inlet temperature to the CGC, and are disc ussed in the Mixing Juncti on section later. The CGC exit temperature was acquire d with a J-type thermocouple. The CGC heat rate was estimated with effectively the same equation as that used for the WGC, as shown in Equation 5.20: CGC fg CGC O H CGC RCV CGCh m h m Q, 2 (5.20) where CGC O Hm, 2 represents the mass flow rate of wate r extracted from the CGC. Again in Equation 5.20 hfg appears, but has been recalculated at this point for both the reduced mole fraction of water vapor and gas pres sure at this state point. The effectiveness calculations performed for this heat exchanger were also identical to those of the WGC, shown above in Equations 5.18 and 5.19. Ducting Section 4-5 This ducting component was defined by th e ducting between the HPC exit thermocouple and High Pressure Recuperator (HPR) inlet thermocouples. A heat balance was deemed necessary for this ducting sect ion due to the placement of th e HPC exit thermocouple. The placement of this thermocouple by the OEM result ed in the measurement of HPC exit gases in the middle of an incidental heating process. Upon emerging from the diffuser, the HPC exit gases passed through an annular passage over ho t, HPT inlet ducting before its egression from the engine. The result of this was significan t heating between the two thermocouples defining this ducting section. This heati ng rate was computing using Equation 5.21: 5 4 5 4 h m QHPC (5.21) where HPCm was calculated by applying the conservati on of mass principle to the HPRTE, per Equation 5.22. CGC O H RCV LPC MAI HPCm m m m m, 2 (5.22)

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60 High Pressure Compressor (HPC) As mentioned above, there was si gnificant heating of the gases leaving the HPC. One of the initial obstacles in charac terizing the HPC performance was to determine the HPC adiabatic exit temperature. With the HPC exit thermocoupl e residing in the middle of this inadvertent heating, it was necessary to uti lize some of the engine manufacture data on the HPC. Included within the engine documentation [16] was a uniquely formatted compressor map for the HPC that served this purpose. This map is shown in Figure 5-1, and provides the adiabatic HPC exit temperature as a function of corrected flow rate and corrected speed. The HPC inlet and adiabatic exit temperatures were then used to find its h, and in turn its work rate, as shown below. HPC HPC HPCh m W (5.23) The HPC isentropic efficiency, HPC, was computed with the same equation as LPC, shown in Equation 5.11. Similar to ducting section 4-5, a heat rate wa s defined to account for the temperature rise between the one-dimensionalized, adiabatic HPC ex it and the HPC exit thermocouple. This heat rate, shown in Equation 5.24 was arbitrarily grouped am ong other HPC parameters and considered as a heating process immediately subs equent to compression, in order to maintain continuity of the flow path. The enth alpy change in this heating process ( h4) was computed using the adiabatic HPC exit temperature found via Figure 5-1, and the temperature acquired by the HPC exit thermocouple. 4 4h m QHPC (5.24)

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61 Combustor The combustor presented several challenges during the formation of this data reduction algorithm. Principle among these was the lessthan-ideal location of the combustor exit thermocouple. This thermocouple was situated about 14 inches (36 cm) from the flame, a radiation source approaching 3500 F (2200 K), and only partway through the mixing of hot primary-zone gases and cool dilution gases. Mo reover, for reasons discussed later, the HPT exit temperature was also difficult to ascertain, making it difficult and unreliable to back calculate the HPT inlet and combustor exit temperatures with a work balance. Thus, a different approach was adopted to estimate the combustor exit temperature as accurately as possibl e using the fuel flow rate, and combustor inlet conditions. Initial efforts approximat ed the combustor as a heater with no losses to ambient, and later evolved to estimat e convective and radiative losses to ambient. The rate of heat added to the working fluid in the combustor was estimated using Equation 5.25: RAD CONV COMB FUEL COMBQ Q LHV m Q (5.25) where LHV is the fuel lower heating value, CONVQ and RADQ are the estimated convective and radiative rates of heat loss from the combustor to ambient. The type of fuel used for these experiments was diesel-2, and its LHV was assu med to be 18,300 Btu/lbm (42.6 MJ/kg). The combustion efficiency, COMB, was defined using the CO and CO2 mole fractions (yCO and yCO2) obtained by the gas analyzer, a nd rests on the assumption that the exhaust gas composition was frozen between the combustor exit and the exhaust stack where gas samples were measured. 21CO CO COMBy y (5.26) To estimate the two heat losses in Equation 5.25 the combustor was discretized into several sections defined by its geometric features. Du ring engine runs, a surface probe was used to

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62 obtain surface temperatures on the combustor. Equations 5.27 show how each of the heat losses were computed: AMB i SURF i i CONVT T A h Q (5.27a) 4 4 AMB i SURF i i RADT T A Q (5.27b) where in each case the losses from each discrete area (Ai) and its corresp onding surface temperature (TSURF,i) are summed over the surface of the combustor. TAMB represents both the temperature of the ambient air and th at of the radiative heat sink, and represents the StefanBoltzmann constant. In Equation 5.27b, the radiative heat exchange is approximated as that between two black bodies, and as such the emissivity, is one. In Equation 5.27a, h is an assumed convection coefficient of about 25 W/m2-K. Incropera and DeWitt [15, p. 8] suggest that this value is at the high end of the natu ral convection regime, and at the low end of the forced convection regime. The actual heat tr ansfer environment ar ound the combustor had a very low velocity flow passing over it, induced by the facility fans, and as such was judged as forced-free convection. In any case, the estimations obtained from Equations 5.27 were consistently less than 1% of QCOMB. Once the heat addition rate was computed from Equation 5.25, the enthalpy change of the working fluid in the combustor, hCOMB, was computed with Equation 5.28: COMB COMB COMBm Q h (5.28) where COMBm is simply the sum of HPCm and FUELm From this point, given the combustor inlet temperature, a subroutine was employed to iterate on the exit temperature until h equaled hCOMB. The iterative process was made necessary by the quartic polynomials, in temperature, that were used to compute h (Equation 5.5).

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63 The equivalence ratio ( ) was calculated using Equation 5.29: STOICHFA FA (5.29) where FA is the actual fuel-to-air ratio, and FASTOICH is the stoichiometric fuel-to-air ratio. The fuel-to-air ratio was computed with Equation 5.30. LPC MAI FUELm m m FA (5.30) It is important to note, as pointed out by Crittende n [17], that the calculations shown in Equations 5.29 and 5.30 fail to consider unburned oxygen cont ributions from the recirculation gases. Therefore, the com bustion environment would actually be leaner than Equation 5.29 would imply, if there is recirculation flow. Ducting Section 7-8 This section of ducting exists between the combustor exit and the HPT inlet, and is characterized by its heat loss to gases exiting th e HPC (Ducting Section 4-5). In a sense, these two segments of the HPRTE flow path can be vi ewed as an internal heat exchanger. The assumption is made both here and for Ducting Sec tion 4-5 that this hypoth etical heat exchanger is well insulated from ambient, si nce it is the cold-side that shar es a surface with ambient air. Proceeding from this assumption, the heat lo st by ducting section 7-8 is equated to the cumulative heat gained by the HPC exit gases be fore reaching the recuperator, per Equation 5.31. 5 4 4 8 7 Q Q Q (5.31) The next step was to use this heat rate to find the enthalpy change, h7-8, COMBm Q h 8 7 8 7 (5.32)

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64 and iterate to find its corresponding HPT inlet te mperature, with the previously calculated combustor exit temperature. High Pressure Turbine (HPT) As was previously mentioned, there were signi ficant uncertainties in directly acquiring the HPT exit temperature. The OEM equipped the HPT exit with four compression fittings to accommodate four HPT exit thermocouples. These are situated around the circumference of the exhaust plane, 90 degrees apart. Specifically, the uncertainties arose fr om the apparent, highly stratified flow exiting the HPT. Each of the four thermocouples consistently indicated four, distinctly different temperatures differing by as much as 320 F (178 K). Rotation of the thermocouples from one location to another conf irmed that the phenomenon was particular to the temperature field around the HPT exit plane, a nd not the thermocouples themselves. It was believed that simply averaging the four temper atures would be a poor representation of the mixed, one-dimensional HPT exit temperature, si nce initial efforts to use a linear average resulted in erroneously high isentropic efficiencies ( HPT > 100%). Consequently, a different approach was employed, seeing as enough info rmation was available from elsewhere to complete the HPT analysis. Instead of attempting to directly calcul ate the HPT power by means of the fluid h, a power balance was implemented, as shown in Equation 5.33: LOSS DYNO HPC HPTW W W W (5.33) where DYNOW is the power absorbed and m easured by the dynamometer, and LOSSW represents the power consumed by friction a nd pumps within the engine. The power losses were provided by the OEM [16] for each of com ponents internal to the core engine: fuel pump, oil pump, bearing friction, and rotor friction. In addition, the total power loss data was conveniently added

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65 by the manufacturer. All of these losses are show n as a function of rotor speed, and can be seen in Figure 5-2. A built-in Matlab function was utilized to interpolate from the total power loss curve using a cubic spline interpolation method. The dynamometer power was calculated using meas urements acquired from a load cell that countered the engine torque. For convenience, a straightforwar d equation was provided [16] by the dynamometer manufacturer, as shown below: 4500DYNO DYNO DYNOF W (5.34) where FDYNO and DYNO are the force (in pounds) acquired by the load cell, and dynamometer speed (in rpm), respectively. The number in the denominator is the dynamometer constant, and accounts for the length of the dynamometer moment arm, and also for unit conversions such that DYNOW is in units of horsepower. Having found the HPT power, and with knowledg e of the HPT flow rate, the enthalpy change across the HPT ( hHPT) could then be com puted with Equation 5.35. COMB HPT HPTm W h (5.35) Again with the component inlet temperature as a starting point, an itera tive subroutine was used with hHPT to find the HPT exit temperature. Upon cal culation of the HPT exit temperature, the HPT isentropic efficiency was computed using Equation 5.36: HPT HPTHPT HPTI HPTX HPTPR T T 11 1 (5.36) where THPTI and THPTX are the HPT inlet and exit absolute temperatures, respectively. The

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66 variable, HPT represents the average specific heat ratio across the HPT, and PRHPT is the HPT pressure ratio defined in Equation 5.37. ATM HPTI ATM HPTX HPTP P P P PR (5.37) Recuperator For most of the Build-4 engine runs the recuperator was well insulated with a two-inch (5 cm) thick, ceramic blanket on top of another thin fiberglass wrap. This effort was primarily made to simplify the data reduc tion algorithm, reducing or eliminating the difficult-to-ascertain heat loss to ambient from the recuperator. The surface temperature of the insulation was measured during engine runs, and a rudimentary an alysis confirmed that losses to ambient were typically three to four percent of the heat transfer rate betw een the hot and cold streams. Nonetheless, the estimation of heat losses through the insulation was retained in the analysis. For engine runs with the insula tion in place, the hot-side heat rate calculati on is shown in Equation 5.38 using the cold-side heat rate and the abovementioned heat rate to ambient: AMB R COLD R HOT RQ Q Q, (5.38) where AMB RQ, represents the heat losses to ambient, and COLD RQ, is the recuperator cold-side heat rate, computed with Equation 5.39, COLD R HPC COLD Rh m Q, (5.39) where hR,COLD was calculated with the two well-placed thermocouples bounding the cold-side of the recuperator. Similar to the combus tor heat losses to ambient, the recuperator AMB RQ, was calculated by breaking the recuperator into thr ee discrete areas and surface temperatures, and then applying Equation 5.40.

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67 3 1 i i i AMB RT A h Q (5.40) The enthalpy change across the recuperator hot side was then calculated using Equation 5.41, COMB HOT R HOT Rm Q h (5.41) which in turn enabled the recuperator hot-side in let temperature (also the HPT exit temperature) to be re-calculated using the recuperator hot -side exit temperature. This new HPT exit temperature was compared to the previously ca lculated value to make sure they corresponded satisfactorily. From this point, the HPT exit temperature was used with hHPT to again make redundant calculations of HPT inlet and combustor exit temperatures and ensure agreement with the previously calculated values. For engine runs that took place before the insulation was utilized a slightly different method was employed. Due primarily to the la rge variations in th e recuperator surface temperature, the heat rate to ambient was estim ated with much less confidence. Additionally, this heat loss was considerably larger, and as such its uncertainty would weigh more heavily on the uncertainty of the hot-side heat rate. For th is reason, redundant calc ulations were not made in these cases, and the recuperator hot-side heat rate was computed with Equation 5.42, HOT R COMB HOT Rh m Q, (5.42) where in this case hR,HOT was found using the previously ca lculated HPT exit temperature and the measured hot-side recuperator exit temperature. The ambient heat losses for the uninsulated cases were found by summing the hot-side and cold-side heat rates. Lastly, the recuperator effectiven ess was computed using Equation 5.43. MAX R COLD RQ Q, (5.43)

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68 Due to its lower average temperature and specifi c heat, the cold stream consistently had the lower heat capacity rate. Therefore, the cold-s ide stream is the only one that can achieve the maximum h, and the maximum heat transfer rate is defined in Equation 5.44, MAX R HPC MAX Rh m Q, (5.44) where hR,MAX was found using the recuperator cold-side and hot-side inlet temperatures. Low Pressure Turbine (LPT) No flow rate measurements were practical given the ducting arrangement of the LPT, only pressures and temperatures at the LPT inlet and exit were acquired. However, a power balance with the LPC made a LPT flow rate estimate possible. This was accomplished by assuming a mechanical energy conversion efficiency, MECH, of 0.97 between the two, allowing the LPT work rate to first be esti mated, as shown in Equation 5.45. MECH LPC LPTW W (5.45) After calculation of the enth alpy change across the LPT ( hLPT) the LPT flow rate was estimated using Equation 5.46. LPT LPT LPTh W m (5.46) A LPT map plotting corrected flow rate over LPT pressure ratio wa s also available to provide a redundant flow rate estimate. The two values consistently agreed sa tisfactorily with one another. The LPT isentropic efficiency was calculated using Equation 5.36 with LPT pressures and temperatures substituted. Mixing Junctions There were several mixing junctions in the HPRTE flow path. Only one of these, however, required an energy balance to find the mixed exit temperature. This mixing junction

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69 existed just upstream of the CGC, where the LPC exit gases mix with reci rculation gases issuing from the WGC. In this case, the mixing proce ss was treated as adiabatic. Furthermore, the temperatures and specific heats of the two inle t streams were close enough to one another to justify the approximation of cons tant specific heats. Used here only, this approximation permitted the temperature of the exit stream to be straightforwardly found using Equation 5.47, which presents the calculation us ing generalized variables. 2 1 2 1 2 2 1 1 1 32P P P Pc c m m T c m T c m T (5.47) In Equation 5.47, the subscript, 3, denotes a variable pe rtaining to the mixed, exit stream, and the subscripts 1 and 2 denote variable s pertaining to the two inlet str eams. The specific heat of the mixed, exit stream is approximated as the linear aver age of the two inlet streams specific heats. Changes in gas composition were also track ed across mixing junctions. It should be pointed out that water extracti on locations were also classifi ed as mixing junctions. Although there was no mixing taking place, per se, a sort of un-mixing of fluids was occurring. Inclusion of water extraction locations with the more typical mixing junctions allowed use of the same equations, generalized for gas composition bookkeeping. The mass flow rate leaving a mixing junction was plainly found with Equation 5.48: 2 1 3m m m (5.48) where the sign of the second mass flow rate is de termined by whether the fluid stream is entering or exiting the mixing junction. Where in Equation 5.47 the subscripts 1 and 2 denoted the two inlet streams, they shall he nceforth denote the two known streams, since information about these two streams is known, and information pertaining to stream 3 is unknown. Next, the molecular weights of the two known streams (M1,2) were computed using Equation 5.2b and the mole flow rates of all three were found using Equations 5.49:

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70 2 1 2 1 2 1M m n (5.49a) 2 1 3n n n (5.49b) where again, in Equation 5.49b, the sign of the second known stream depends whether its mixing with stream 1, or exiting the flow path. From here, the mole flow rates of each stream were multiplied with their corresponding vector of mole fractions to obtain mole flow rates of each pure substance within each of the known streams, as shown below. 2 1 2 2 2 2 2 1 2 1 2 2 2 2 CO O H CO O N CO O H CO O Ny y y y y n n n n n n (5.50) Finally, the mole flow rates from the known stream s were used to find the mole fraction vector of the unknown stream, as seen in Equation 5.51. 2 2 2 2 2 3 1 2 2 2 2 3 3 2 2 2 21 1 CO O H CO O N CO O H CO O N CO O H CO O Nn n n n n n n n n n n n y y y y y (5.51) In the special case of water extraction locations, the sign in front of stream 2 was negative, and the mole fraction vector elements of stream 2 we re all zero, except for the fourth element, which was one. Propagation of Uncertainty In addition to reducing and in terpreting the experimental data, a thorough propagation of uncertainty analysis was performed. To comple te the uncertainty analysis, classical methods were employed as presented by Holman [18]. For a given dependent parameter, R, with n

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71 independent parameters as shown in Equation 5.52, nx x x f R, ,2 1 (5.52) the uncertainty of R was computed using Equation 5.53, 2 1 2 2 2 2 2 2 2 1 2 1 Xn n X X Rx R x R x R (5.53) where Xn is the uncertainty of the nth independent parameter. Equation 5.53 was applied to find the uncertainty of all the parameters discussed in the above section. All uncertainties of reduced parameters stem med from one of three possible sources: instrumentation error, data acquisition error, and miscellaneous transient behavior, either real or artifact. This concept is generalized by Equation 5.54. TRAN X DAQ X INST X X (5.54) Ideally, all three sources would be comparable and small. However, it is believed that the three contributions usually differ ed considerably, and were most frequently dominated by the third term. Typically, data were collected over a period ranging from five minutes to 15 minutes, and during this time nearly all of the temperatures, pressures, and flow rate s either climbed, fell, or fluctuated to some degree. Since the objec tive was to characterize a single, steady operating point by averaging several hundred data points, many of the uncerta inties were often taken as one standard deviation from the mean. For some parameters, such as ambient temperature, there was very little fluctuation aside from some noise in the signal. In such cases, the uncertainty in the instrument itself was the dominating source of error. Error derived from the AD conversion of signals was considerably smaller than the other two sources. 16-bit resolution was used within the DAQ system to curb quantization errors, and the time between samples was sufficiently large to avoid any appreciable aperture errors.

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72 Occasional testing was also carried out to spot-c heck for significant non-linearity errors made by the AD conversion. In general the uncertainty of all the measured values was taken to be the greater of the instrument error and one standard deviation from the mean value. Only one standard deviation was thought necessary because the data were typically skewed by relatively few outliers. These outliers could result from several occurrences, e. g. the momentary loss of a thermocouple signal, or indeterminate electromagnetic interference. The dynamometer water supply pressure would also fluctuate erratically due to other plumbing loads in the building. This often caused brief excursions in the engine load, engine speed, fuel flow rate, and other parameters as well. Instrument uncertainties for thermocouples, pr essure transducers, and other instruments can be found in Table 5-1. Data Adjustments The reduced data described in the above sect ions were obtained from a very particular engine and combination of compone nts. Consequently, some of th e capabilities of this combined cycle were veiled by excessive pressure drops, poor VARS performance, and high ambient/PCW temperatures. In addition, comparisons between these experimental results and other experiments or models needed to be made on the same basis. For these reasons, classical, partially non-dimensional parameters were used to present the data in a properly scaled form. The corrected parameters chosen are explained in greater detail by Volponi [19], and are shown in Table 5-2. As mentioned above, the pressure drops in the HPRTE ducting was high enough to obscure the experimental results. Specifically, when transitioning into the pressurized regime, the HPT was initially back-pressured in order to accelerate the LPT. This resulted in a premature increase in the HPT exit and inlet temperatures, such that ve ry little or no margin was left for loading the

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73 engine by the time the LPC exit pressure rose ap preciably. Furthermore, various leaks in the ducting and engine casing intensified with increas ing boost due to the incr eased pressure within, resulting in an un-metered loss of working fluid. Consequently, nearly all of the pressurized data sets were at or near maximum HPT inlet temp eratures, and at reduced engine speeds with no load on the dynamometer, resulting in an uninfor mative scatter of net power around zero. This necessitated a supplementary mode ling approach to approximate th e net power if the pressure drops had been different, since in reality th is can be an arbitrary design choice. Pressure Drop Considerations In the interest of better evaluating the HP RTE performance, several adjustments were made to the raw data, and a supplementary mode ling approach was employed. In this modeling approach, the empirical pressure drops in the reci rculation flow path and within the recuperated engine. This provided an upper-bound estimate of HPRTE power and efficiency in the absence of any pressure drops. The first pressure ad justment reduced the hot -side recuperator and recirculation pressure drops to zero by setting the HPT exit pressure equal to the HPC inlet pressure. The second adjustment eliminated th e core-engine pressure drop by raising the HPT inlet pressure to the HPC exit pressure. The above adjustments translate to increased ne t power in that the HPT pressure ratio is increased, resulting in a greater temperature ra tio and temperature difference. The new HPT expansions began from the same temperatures, but were completed at lower temperatures. The lower HPT exit temperatures were calculated by a ssuming the HPT efficiency to be constant in all cases. Though no maps were available for this turbine, this was judged as a fair approximation when considering a ma p generated by AXOD [20] in Figure 5-3. There it can be seen that the HPT efficiency is approximately 85% for a wide range of corrected speeds and pressure ratios bounding the rang e of values seen in experiments. The adjusted HPT exit

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74 temperature was computed using Equation 5.55, HPT HPTHPTI HPTX HPT HPTI HPTXP P T T 11 1 (5.55) where HPT is 0.85. Having estimated the adjusted HP T exit temperature, the adjusted net work output from the HPRTE was computed using Equation 5.56. HPC HPT COMB NETW h m W (5.56) In the interest of further generalizing the reduced results, it can also be seen from Equation 5.56 that the friction work has been ignored in this computation. The HPC work term, HPCW was calculated using Equation 5.23. Table 5-1. Uncertainty of instruments Instrument Uncertainty T-Type Thermocouples Greatest of 0.9 F (0.5 K) or 0.4% J-Type Thermocouples Greatest of 2.0 F (1.1 K) or 0.4% K-Type Thermocouples Greatest of 2.0 F (1.1 K)or 0.4% 0-30 Pressure Transducers 0.1% of F.S. (typical), 0.5% (max) 0-100 Pressure Transducers 0.1% of F.S. (typical), 0.5% (max) Differential Pressure Transducers 0. 1% of F.S. (typical), 0.5% (max) Chilled Water Flowmeters 0.5% Fuel Flowmeter 1.0% Water Extraction Load Cell 0.2 % of F.S. Dynamometer Load Cell Unknown Engine Speed Optical Tachometer 3 rpm

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75 Instrument Uncertainty Gas Analyzer UBHC: 20 ppm CO: 0.03% CO2: 0.6% O2: 0.2% NO: 10 ppm Table 5-2. Data scalin g parameters [19]. Parameter Corrected Parameter Rotor Speed N N Gas Mass Flow Rate m m Power W W Temperature T T Pressure P P

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76 Figure 5-1. Map used for finding the HPC adiaba tic exit temperature, shown as a function of corrected flow rate and corrected speed [16].

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77 Figure 5-2. Rover 1S-60 internal power lo sses as a function of rotor speed [16].

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78 Rover Turbine Efficiency Map0.5 0.55 0.6 0.65 0.7 0.75 0.8 0.85 0.9 0.95 11.522.533.54Pressure RatioEfficiency 615.7 N/sqrt(T4) 718.3 1128. 7 1026. 1 923.5 820.9 Figure 5-3. Rover 1S-60 turbin e map generated by AXOD [20].

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79 CHAPTER 6 EXPERIMENTAL RESULTS The experimental apparatus described in Chapter 3 was operated according to the procedures discussed in Chapter 4 on 13 occasions to yield 34 distinct sets of working data. The data reduction algorithm outlined in Chapter 5 wa s then applied to the data, providing the relevant performance parameters. As was mentioned in Chapter 5, all but five of these data sets were idle runs, this was due primarily to exces sive pressure drops and a slightly mismatched turbocharger. Since no load was applied to th e dynamometer, the resulting dataspecifically the net power and efficiencywere less informative, as shown in Figure 6-1. Since the engine was in a sense loaded by back-pressuring the HP T, the steps outlined at the end of Chapter 5 were taken to estimate the power consumed by the various loss mechanisms and present the adjusted power. All of the adjusted, reduced da ta generated by the methods discussed in Chapter 5 are tabulated in Appendix C. Worth noting is th at all of the data sets were at or near the maximum sustainable HPT inlet temperature, and as such represent full power engine runs, while other operating parameters such as recirculation ratio and LPC pressure ratio were varied. For conciseness the values previously referred to as adjusted power will henceforth be referred to as power, since it is the only power of interest. Gas turbine engines exhibit a particularly strong dependence on ambient conditions, namely pressure, temperature, and humidity. However the HPRTE is unique in that the inlet pressure of the core engine can be changed by co ntrolling the LPC pressure ratio. Furthermore, when coupled with the VARS, the HPC inlet temper ature can also, in principle, be changed at the will of designers and ope rators. For these reasons, and the classical correction parameters defined in Table 5-2, were calculated using the HP C inlet conditions instead of the

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80 ambient conditions, since ambient conditions possess little influence over the HPRTE performance. Although most turbine engines, including th e Rover 1S-60, have the capability of operating at various speeds, no atte mpts were made in these experiments to capture data at offdesign speeds. However, as the HPT was back-p ressured and the turbocharger accelerated, there was significant speed-droop resulting from the m echanical governor, which helps to validate the idea that the engine was, in effect, being loaded Additionally, the correc ted speed (N*), defined in Table 5-2 showed even greater variance, since the HPC inlet temperature fluctuated as well. Since N* refers performance parameters to standard ambient conditions, it was used as a dependent parameter in place of engine speed (N ). In the interest of characterizing the performance of this engine, several plots were generated. The first of these is Figure 6-2, which plots corrected power (P*) and thermal efficien cy against N*. Also included within Figure 6-2 is the adjusted power before inlet-condition correcti ons (P). A peculiar characteristic of the P and P* data is that they have opposite slopes when plotted over N*. This results from the fact that the scaling parameters generally demonstrated a dependence on engine speed. Ideally, any range of HPC inlet pressures and temperatures could be actualized at any ope rating point (any N* for example), but due to the operati onal boundaries of the engine a nd VARS, these inlet conditions were coupled with engine speed. The scaling parameters, and are shown over N* in Figure 6-3 for comparison with Figure 6-2. I can be seen in Figure 6-3 that both scaling parameters increase when decreasing N*. In the case of this trend is more straightforwar d. The higher LPC pressure ratios (essentially the same values as ) were attained only at lower engine speeds, because increased operation of the turbocharger directly caused the speed-droop, as described in the Data

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81 Adjustments section of Chapter 5. The trend also appears to follow speed, but more directly correlates with LPC pressure ratio, with some random fluctuations as well. Again, as the turbocharger was accelerated, great er flow rates were diverted from the exhaust and driven through the LPT and through the recirculation flow path. In addition, as the HPT back-pressure initially rose, the HPT pressure ratio decrea sed until the LPC could be accelerated enough to compensate, increasing the HPC inlet pressure. This caused the exhaust gas temperature (EGT) to climb. As the engine speed dropped, the HPT pressure ratio continued to fall as well, further increasing EGT. These two consequences of pr essurizationincreased recirculation flow and temperaturecombined with one another to greatly increase the heat load on the VARS generator (the HGC) in particular, but on the WGC and CGC also. The resulting heat loads taxed the VARS and the WGC exceedingly and alwa ys effected an increase in HPC inlet temperatures. Thus, tended to scale with LPC pressure ratio, and in turn speed. The relationship between and LPC pressure ratio is shown in Figure 6-4. When considering the trends in Figure 6-3, it is important to point out that itself appears in the N* parameter as shown in Table 5-2. As a result, the obvious decrease in N* when decreasing N is compounded by the related upsurge in HPC inlet temperature. After considering the trends of the correction parameters re lative to N*, the opposite slopes of P and P* can be better understood. Since both and appear in the denominator of P*, and both have downward slopes over N*, it is more r easonable to expect the results shown in Figure 6-2 provided the denominator of P* ( ) changes faster than its numerator (P). In general, P would be expected to scale one-to-one with HPC inlet pressure such that a given increase in would be accompanied by an equivalent increase in P. However, competing effects somewhat

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82 mitigated the desired power-pressure relationship. Power is plotted against LPC pressure ratio in Figure 6-5. Figure 6-5 shows the power obtained from the adjusted data, as well as the power predictions of a model developed by Khan [8, 12] The model is a steady-state, one-dimensional thermodynamic model of the HPRTE/VARS. The adjusted experimental power and modeled power differ significantly, as seen in Figure 6-5. An initial, sharp rise in P is observed over LPC pressure ratio, but a weaker dependence at higher pressure ratios. The modeled results, however, predict a steady, linear increase in power over LP C pressure ratio. The magnitude of power predicted by the model is most likely lower due to the inclusion of frictio n, parasitic horsepower, and pressure drops by Khan, in an effort to most accurately represent this engine and its configuration. The adjusted power, shown in blue, was an upper-bound estimate using neither pressure drops nor friction power. The anticipated ri se in P is more directly related to the engine mass flow rate, which itself depends on the HPC inlet pressure assuming the inlet temperature is held fast by the VARS. In actuality the inlet te mperatures crept up with the higher LPC pressure ratios (Figure 6-4), lessening the density and potential mass flow rate, although this was a weak effect. A more prevailing de traction from the potential flow rate was the speed-droop also associated with higher inlet pressures (Figure 6-3). The deceleration of both the HPC and HPT caused a diminution in the flow rate through, and pressure ratio across each. With the reduced pressure ratios came also the reduction of temp erature ratios, and therefore smaller temperature differences across the high pressure turbomachinery. In consequence, while the mass flow rate did manage to rise alongside of the LPC pr essure ratio, the HPC and HPT temperature differences dropped. Figure 6-6 illustrates these opposing trends in a dimensional sense, while Figure 6-7 shows the same data, but normalized. The mass flow rate data in Figure 6-7 are

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83 normalized by their minimum, and the delta-T data by their maximum. In this way the percent changes in each of these parameters can be view ed, and the slopes compared. It can be seen in Figure 6-7 that despite the abovementioned mitigating factors, the percent increase in mass flow rate was greater than the decline in turbomachinery temperature differences. The manner in which these parameters influence power can be seen in Equations 5.56 and 5.23. Another suspicion that can be verified in Figure 6-7 is that the mass flow rate increase was less than expected. This is realized by noting that the sl ope of percent-change-inflow-rate regression line indicates only an 84% increase in flow rate, per unit increase in LPC pressure ratio. Again, with ideal hardware and control schemes, an a pproximate 100% increase in flow per unit LPC pressure ratio increase is expected. The engine mass flow rate and corrected ma ss flow rate are shown over N* in Figure 6-8 for comparison with Figure 6-2. The trends in Figure 6-8 are similar to those in Figure 6-2 in that the uncorrected values bear a downward sl ope and the corrected values an upward slope. This comparable behavior can in like fashion be explained by the relationship and exhibit with N*. One difference to be noted is the relati vely gentle slope of corr ected flow rate over N*, where P* demonstrated a much closer relationship with N*. However, 0.5 appears in the numerator of the corrected flow rate parameter, and in the denominator of the corrected power parameter. Thus the downward trend of with increasing N* serves to somewhat lessen the impact of instead of augmenting it, as was so for P*. Thus a greater mass flow rate (and ) at lower speeds was more than balanced by the greater HPC inlet pressure. Having considered the dependence upon inlet conditions and mass flow rate, other nondimensional parameters and their effects on power were also explored. In particular, two cycle parameters were thought to have a significant bearing on power: the HPC pressure ratio, and the

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84 overall Cycle Temperature Ratio (CTR). The CT R is defined as the ratio of the maximum, absolute cycle temperature (HPT inlet temperat ure) to minimum absolute cycle temperature (HPC inlet). These parameters influence can most easily be understood by examining a temperature-entropy (T-S) diagram with an ideal Brayton cycle, shown in Figure 6-9. The diagram in Figure 6-9 shows all temperatures normalized to the inlet temperature such that CTR appears on the ordinate. Three diffe rent cycles are shown on the T-S plot for comparison: a baseline cycle, a higher CTR cycl e at the same pressure ratio as baseline, and a higher HPC pressure ratio at the same CTR as base line. The baseline cycle follows the path 1-23-4, and the power-per-unit-mass is equivalent to the area contained by the cycle, areas A plus B. The second cycle follows the path 1-2-5-6, and clearly yields a signifi cantly higher specific power with areas A, B, and C. The third cycle fo llows the path 1-7-8-9. This cycle differs in that it adds to the specific power with area D, but subtracts from sp ecific power by surrendering area B. In general, there will be a net gain in specific power, but a cycle having different proportions when sketched on a T-S diagram coul d conceivably break even or lose specific power. For this reason, specific power is expect ed to scale more strongly with CTR than with HPC pressure ratio, as is made evident by the data shown in Figure 6-10. Figure 6-10 shows specific power over CTR, but the data are grouped according to their HPC pressure ratios so that the influen ce of each cycle parameter can be examined independently. Regression lines were added only to better illustrate trends made less distinct by the data-scatter. Several of the lines are fit to only three to four data points, and their slopes can be misleading. The data in Figure 6-10 clearly show the closer relationship between CTR and specific power. Additionally, at a constant CT R, specific power climbs somewhat with HPC pressure ratio, but to a lesser, even dimini shing extent at higher pressure ratios.

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85 In Figure 6-11, P* is also presented over CTR for va rious HPC pressure ratios. The data in Figure 6-11 indicate a somewhat stro nger influence of HPC pressure ratio on P* than of CTR on P*. This is most likely related to the closer dependence of HPC pressure ratio on N*. HPC pressure ratio followed N* very closely, wh ere CTR had much less dependence thereupon, as shown by Figure 6-12. Though, incidentally, the HPC in let temperature (the denominator in CTR) did tend to follow speed, that relationship is not as strongly coupled as that of HPC pressure ratio and N*. Furthermore, the HPT in let temperature had no reas on to trend with speed since it is an independently controlled variable, and all of these data sets were maintained at or near the engine maximum HPT inlet temperature. The cycle thermal efficiency may also be exp ected to behave similarly relative to CTR and HPC pressure ratio. The efficiency data exhib it more scatter than the power data, and for this reason two separate figures are used to show their dependence on CTR and HPC pressure ratio. Figure 6-13 shows efficiency over CTR and Figure 6-14 plots it over HPC pressure ratio. In both charts the thermal efficiency demonstrates to a small extent the anticipated, upward trend with CTR and HPC pressure ratio. Though these da ta are more widely dispersed, the stronger correlation with CTR can again be discerned. It is more likely that the ostensible, direct relationship between HPC pressure ratio and ther mal efficiency is artif act, since the higher pressure ratios happened to coincide with th e higher CTRs. More importantly, the higher pressure ratios also occurred together with higher corrected speeds. Figure 6-15 shows that the HPC isentropic efficiency increased with N* as it approached the design speed of 46,000 rpm. The more efficient HPC operation probably contri buted the most to any apparent relationship between overall thermal efficiency and HPC pre ssure ratio. Any similar effects from the HPT efficiency are lost, since it was fixed at 0.85 fo r the power adjustment calculations. Overall,

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86 there was little variation in ther mal efficiency, largely because all of the data points are at fullpower. A principal dimensionless parameter associated with the HPRTE is the recirculation ratio (R) defined by Equation 5.13. This parameter indicates the extent to which the engine is operating within the semi-closed regime. It is generally controlled by changing the LPC pressure ratio, but also depends on the differential pres sure between the hot-side recuperator exit and HPC inlet state points. In a physical sense, this is th e pressure drop of the recirculation circuit. The Build 4 engine configuration was equipped with a va lve in the recirculation flow path with which to vary this differential pressure. This valve, shown as VREC in Figure 1-1, proved essential in making transitions into the pressurized regime and providing a finer control over recirculation flow rate, and over R. Figure 6-16 shows the recirculation ratios against LPC pressure ratio. Again, the modeled results are included for comparison. The experimental data in Figure 6-16 are discretized into two distinct bands. This stems from the neces sity to throttle the recirculation valve for operational reasons. Specifically, as the turbocharger was brought online by partially closing the exhaust valve (VEXH in Figure 1-1), the recirculation flow rate would climb much more quickly than the LPT flow rate. Left unchecked, this wo uld result in excessively high recirculation ratios for the combustor being used, and subsequent en gine shut-downs. Thus, restriction of the recirculation flow was essential to keeping the en gine running stably unti l the turbocharger came up to speed. Once this was accomplished, the tran sition from partially closed to fully open was promptly executed. For data sets in which VREC was about 50% open or less, the recirculation line was considered as restricted, while for all other data sets, it was considered unrestricted. This discretization of data into only two gr oups for a valve with co ntinuous operation was

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87 appropriate, as this particular valve was very seldom between half-open and full-open. The two trends exhibit the same shape as each other, but are shifted apart. This is due to the dependence of R upon the pressure drop through the recirculati on line, since the restri cted values (higher pressure drop) are translated downward from the unrestricted values (lower pressure drop). The experimental trends resemble the modeled tr end, though they are shifted downward. This indicates that the model simulated the performa nce of a HPRTE having an even less restrictive recirculation line. The higher recirculation flow s resulted in R values that could barely be achieved in practice with the Build 4 combustor. The model also predicted a turnabout in the trend at higher pressure ratios, though with the Build 4 configurat ion, these higher pressure ratios could not be achieved. The experimental results do, howev er, exhibit the same upwardly concave shape as the modeled results, indicating th at they would likely cha nge direction if higher pressure ratios were reached. An important role of semi-closure is to re duce the flow rate of fresh air necessary to operate the cycle. As more of the working fluid is composed of recirculated exhaust gases, the demand for fresh air is decr eased as shown in Figure 6-17. For higher values of R the experimental air flow rates appears to remain somewhat constant. These data correspond with the data sets recorded when moving through the transition into the pressurized regime. As mentioned above, when VEXH was initially closed, the recirculat ion flow rate rose more quickly than the LPC pressure ratio, and in turn air flow rate. Essentially, the data in Figure 6-17 are very similar to the data in Figure 6-16, but with the ordinate and abscissa exchanged. This is also seen in the modeled air flow rate trend. The modeled trend is double-valued when shown over R, pointing to the relationship between LPC fl ow rate and LPC pressure ratio. Again, the

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88 experimental trends have a steeper slope at lower R values, and may also mimic the modeled trend if the Build 4 configura tion could advance further into the pressurized regime. In addition to reducing the air flow rate, semi -closure also serves to reduce the exhaust flow rate of the system as well. Beyond decrea sing the size of the turboc harger, a lower exhaust flow rate will also decrease the heat signature of the HPRTE, which can be paramount in some applications. Since the air flow rate and exhaust flow rate differ very little, the data in Figure 617 can also be used to approximate the exhaust flow rate over R. Semi-closure also changes the HPRTE combustion environment. Since the combustor inlet gases are being diluted with exhaust gase s, the stoichiometry within the reaction zone changes with R at a given total flow rate. Figure 6-18 shows how the equivalence ratio increases with R alone. The experimental trends appear to agree well with results predicted by the model. Though the modeled recirculation ratios were much hi gher, the slopes of all three trends are very similar. The experimental data in Figure 6-18 separate into the two bands, but show some intermingling of the restricted and unrestricted cases at higher R values. This points to the dependence of equivalence ratio on other operating parameters as well. Clearly, in Equation 5.29, the equivalence ratio depend s on the air and fuel flow rates. However, with all of these data sets at full power, this is expected to be influenced more by the air flow rate, which appears in the denominator and was more varied in its values. The equivalence ratio can therefore be related to another, primary opera tional parameter, the LPC pressure ratio, as demonstrated in Figure 6-19. The experimental data in Figure 6-19 tend to deviate less from their respective bands. The experimental and modeled trends be ar a close likeness to the trends in Figure 6-16, which reflects the similar, inverse relationship with air-flow-rate that R and equivalence ratio share.

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89 The more pertinent way in which R affects the combustion environment can be measured in the composition of the exhaust gas. The impa ct of R on emissions is here quantified by the concentrations of NO and CO, the least desirabl e constituents existing in significant amounts. The concentrations of these two gases in the exhaust mixture are shown in Figure 6-20. As expected, and is typically the case, a reducti on of the NO concentration is compromised by a corresponding increase in CO. This results fr om the opposite dependences of their equilibrium concentrations on temperature and O2 concentrations. The formati on rate of NO scales linearly with the product of O2 and N2 concentrations, and with temperature cubed. As R is increased, the N2 concentration remains rela tively unchanged while the O2 concentration drops off, as affirmed by the data in Figure 6-21. The N2 data in Figure 6-21 indicate a slight rise with R, but regression lines were added to help illuminate th e difference in slopes. It is clear that the decrease of O2 concentration over R is faster than the rise in N2 concentration. While this dilution of gases detracts from the conditions fo r NO formation, it also diminishes the rate at which CO can further oxidize to form CO2. Consequently, the CO concentrations can be seen to rise with R. Another influential factor is th e temperature at which th ese reactions are taking place. The effects of lessening the O2 concentration are two-fold in that the flame temperature in the primary zone of the combustor is also reduced This reduction in temperature further serves to suppress NO formation, but again, this effect will also reduce the re action rate to finish oxidizing CO. An unexpected phenomenon is evident in Figure 6-20. At the high end of the R range, the NO concentration takes a marked upward swing. This turnabout cannot be rationalized with any of the above arguments, since, at still higher R values, all of the NO-diminishing effects are increasingly present. The most likely reason is related to a minor deficiency in the HPRTE

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90 hardware. The original rover combustor was mo dified to accommodate recirculation flow by welding straps over some of the dilution holes. This forces more of the oxygen-dilute gases through the combustor primary zone for combustion with the intention of restoring oxygen to the reaction. While this first modification worked we ll for moderate R values (R < 1.3), it proved to be insufficient in redirecting O2 at higher R values. As R was increased, the stoichiometry in the primary zone would progressively become leaner. Instabilities of the flame would then follow, viewed as violent flickering in the combustor vi ew port, until the flame relocated downstream to the dilution holes, where more oxygen was present. After the flame relocated, the geometry of the combustion environment changed, with a grea tly increased time for NO forming reactions to take place. A particular benefit of cooli ng the recirculated exhaust gase s is the opportunity to extract the resulting water condensate. However, before this can be achieved to an appreciable extent, the exhaust recirculation mixture must first be cooled to its dew-point. This was accomplished with the VARS heat exchangers and the WGC, and relied on the performance of each. Unfortunately, the VARS employed for these expe riments was somewhat undersized to provide the refrigeration necessary for respectable water extraction. Another non-dimensional paramete r is introduced to present the water extraction data: the mass flow rate of extracted water, normalized by the fuel mass flow rate. This Water to Fuel Ratio (WFR) is plotted against th e LPC pressure ratio in Figure 6-22. The theoretical maximum for WFR on a volume basis is around one, which, on a mass basis, equates to approximately 1.2, though these values can only be approached at high recirculation ratios. Taking this into consideration, it can be said that the water ex traction performance was respectable at lower LPC pressure ratios. However, the WFR subsequent ly drops as the tests moved further into the

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91 pressurized regime. The modeled results predicted a much higher WFR throughout the pressurized regime. However, th e model also simulated an ideal VARS equipped with sufficient refrigeration capacity to maintain a constant, 59 F (519K) CGC exit temperature. Thus the modeled WFR depended only on the gas proper ties at the CGC inlet, and not on VARS performance. There are two competing effects relating to HP RTE operation that most directly contribute to WFR: R and LPC pressure ratio. Greater concentrations of water in the CGC inlet gas mixture will lead to a greater potential for water extraction by raising the dew-point temperature. The recirculating gas by itself ha s a high water concentration, but is dried considerably when mixed with the fresh air from the LPC. Thus, th e greater values of R can be expected to yield higher WFR values, and R drops with increasing pressure ratio, following the trend seen in Figure 6-22. However, the effect of increasing the pressure in the recirculation line (increasing the LPC pressure ratio) should raise the saturati on pressure of the water vapor along with its dew-point, initiating water extrac tion earlier in the cooling process. Regardless of these two factors contending with one anothe r, the fact remains that the VARS and heat exchangers must perform well for appreciable water extraction to be realized. The overriding factor that governed water extraction performance was undoubtedly the suffering pe rformance of the VARS and WGC at higher LPC pressure ratios. The h eat load on the three recirculation line heat exchangers is related to the LPC pressure ratio by the recirculation flow rate, which was driven upward with LPC pressure ratio by throttling VEXH, as shown by the data in Figure 6-23. The recirculation gas mixture did reach its de w-point in the CGC most of the time, but by this point the refrigeration capacity was usua lly exhausted. Although there were no means available to directly measure the humidity within the saturated state of the gas was made evident

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92 by noteworthy quantities of liquid water finding its way through leak -paths in the ducting of the HPC. Figure 6-24 illustrates how the dew-point temper ature changed with the LPC pressure ratio. Figure 6-24 also shows the increases in CGC inle t and exit temperatures associated with further pressurization of this HPRTE rig. It is apparent that, thought th e dew-point temperature did increase with LPC pressure ratio, the evaporat or temperatures increased at a greater rate. The water extraction rate should roughly scale w ith the difference between the inlet dew-point temperature and CGC exit temperature. Althoug h the CGC temperature difference increased at higher pressure ratios, the CGC exit temperature still climbed qui ckly enough to approach the dew-point temperature, lessening the water extraction rate. 0 2 4 6 8 10 12 14 16 0.911.11.21.31.41.51.6LPC Pressure RatioNet Power (hp) & Efficiency (%) Power Thermal Efficiency Figure 6-1. Actual net power output from Build 4 engine runs.

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93 y = -3E-07x2 + 0.038x 950 R2 = 0.727 y = -5E-06x2 + 0.421x 9176 R2 = 0.462 y = -1E-09x2 + 0.0001x 2.96 R2 = 0.2920 10 20 30 40 50 60 70 80 90 43.043.544.044.545.045.546.0 ThousandsN* (rpm)P* and P (hp)0.11 0.13 0.15 0.17 0.19 0.21Thermal Efficiency P* P Efficiency Figure 6-2. Corrected Power, power, and th ermal efficiency over corrected speed. y = -6E-08x2 + 0.005x 102 R2 = 0.832 y = 2E-09x2 0.0002x + 6.64 R2 = 0.9221.00 1.05 1.10 1.15 1.20 1.25 1.30 1.35 1.40 1.45 1.50 43.043.544.044.545.045.546.0 ThousandsN* (rpm) and Figure 6-3. Scaling parameters versus corrected speed.

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94 y = -0.062x2 + 0.277x + 0.80 R2 = 0.8121.01 1.02 1.03 1.04 1.05 1.06 1.07 1.08 1.09 1.1 1.01.11.21.31.41.51.6LPC Pressure Ratio Figure 6-4. Theta versus LPC pressure ratio. y = -42.2x2 + 129x 23.0 R2 = 0.56130 40 50 60 70 80 1.01.21.41.61.82.0LPC Pressure RatioPower (hp) P Modeled Figure 6-5. Power versus LPC pressure ratio.

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95 y = 18.9x2 100x + 442 R2 = 0.802 y = 19.7x2 65.6x + 302 R2 = 0.676 y = -0.870x2 + 3.27x 1.12 R2 = 0.9470.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8 2.0 1.01.11.21.31.41.5LPC Pressure RatioMass Flow Rate (lbm/s)230 280 330 380 430HPT and HPC T (F) Mass FLow Rate HPT T HPC T Figure 6-6. High pressure tur bomachinery temperature differences and mass flow rate versus LPC pressure ratio. Slope = -14.83 Slope = 84.33 Slope = -6.44-10 0 10 20 30 40 1.01.11.21.31.41.51.6LPC Pressure RatioPercent Change in Flow Rate, HPT T and HPC T Mass Flow Rate HPT T HPC T Figure 6-7. Percent change in high pressure turbomachinery temperature differences and mass flow rate versus LPC pressure ratio.

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96 y = -2E-08x2 + 0.001x 29.8 R2 = 0.243 y = -1E-07x2 + 0.009x 198 R2 = 0.7581.0 1.1 1.2 1.3 1.4 1.5 1.6 1.7 1.8 1.9 43.043.544.044.545.045.546.0 ThousandsN* (rpm)Mass Flow Rates (lbm/s) Corrected Mass Flow Rate Mass Flow Rate Figure 6-8. Flow parameters versus corrected speed. -0.5 0 0.5 1 1.5 2 2.5 3 3.5 4 0.5 1 1.5 2 2.5 3 3.5 4 CTRs/R 1 2 7 3 4 5 6 8 9 A B C D Figure 6-9. Generalized effect s of HPC pressure ratio and CT R on specific power, illustrated on a T-S diagram.

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97 3.3 3.35 3.4 3.45 3.5 3.55 40 41 42 43 44 45 46 47 48 49 50 CTRSpecific Power (hp-s/lbm) 2.60-2.66 2.69-2.72 2.73-2.76 2.81-2.83 2.85-2.87 2.89-2.91 Figure 6-10. Specific power versus CTR, for various HPC pressure ratios. 3.3 3.35 3.4 3.45 3.5 3.55 46 48 50 52 54 56 58 60 62 64 CTRP* (hp) 2.60-2.66 2.69-2.72 2.73-2.76 2.81-2.83 2.85-2.87 2.89-2.91 Figure 6-11. Corrected power versus CT R for various HPC pressure ratios.

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98 y = 1E-08x2 0.0009x + 21.4 R2 = 0.492 y = -2E-09x2 + 0.0004x 8.84 R2 = 0.9582.5 2.7 2.9 3.1 3.3 3.5 3.7 43.043.544.044.545.045.546.0 ThousandsN* (rpm)HPC Pressure Ratio and CTR HPC Pressure Ratio CTR Figure 6-12. HPC pressure ratio and CTR versus corrected speed. y = -0.327x2 + 2.30x 3.91 R2 = 0.6580.120 0.125 0.130 0.135 0.140 0.145 0.150 3.253.303.353.403.453.503.55CTREfficiency Figure 6-13. Thermal efficiency versus CTR.

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99 y = -0.084x2 + 0.505x 0.617 R2 = 0.3490.120 0.125 0.130 0.135 0.140 0.145 0.150 2.552.602.652.702.752.802.852.902.95HPC Pressure RatioEfficiency Figure 6-14. Thermal efficiency versus HPC pressure ratio. y = -1E-07x2 + 0.011x 201 R2 = 0.57669.0 69.5 70.0 70.5 71.0 71.5 72.0 72.5 73.0 43.043.544.044.545.045.546.0 ThousandsN* (rpm)HPC Isentropic Efficiency (%) Figure 6-15. HPC isentropic efficiency versus corrected speed.

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100 y = 0.808x2 2.91x + 3.74 R2 = 0.882 y = 5.38x2 16.8x + 14.4 R2 = 0.9131.1 1.2 1.3 1.4 1.5 1.6 1.7 1.8 1.9 2.0 1.01.21.41.61.82.0LPC Pressure RatioR Restricted Unrestricted Modeled Figure 6-16. Recirculation ratio versus LPC pressure ratio. y = 0.795x2 2.71x + 2.95 R2 = 0.975 y = 1.102x-1.623R2 = 0.9190.40 0.50 0.60 0.70 0.80 0.90 1.00 1.10 1.11.21.31.41.51.61.71.81.92.0RAir Flow Rate (lbm/s) Restricted Unrestricted Modeled Figure 6-17. Fresh air flow rate versus recirculation ratio.

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101 y = 0.358x0.610R2 = 0.643 y = 0.032x2 + 0.213x + 0.089 R2 = 0.9710.35 0.40 0.45 0.50 0.55 0.60 1.11.21.31.41.51.61.71.81.92.0REquivalence Ratio Restricted Unrestricted Modeled Figure 6-18. Equivalence ratio ve rsus recirculation ratio. y = 0.337x2 1.13x + 1.32 R2 = 0.959 y = 0.649x-1.178R2 = 0.6170.35 0.40 0.45 0.50 0.55 0.60 1.01.21.41.61.82.0LPC Pressure RatioEquivalence Ratio Restricted Unrestricted Modeled Figure 6-19. Equivalence ratio versus LPC pressure ratio.

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102 y = 63.2x2 178x + 137 R2 = 0.610 y = 0.280x2 0.087x 0.162 R2 = 0.3590 5 10 15 20 25 30 35 40 45 0.70.91.11.31.51.71.9RNO Concentration (ppm)-0.10 -0.05 0.00 0.05 0.10 0.15 0.20 0.25 0.30 0.35 0.40CO Concentration (%) NO CO Figure 6-20. Nitric oxide and carbon monoxide concentrations ve rsus recirculation ratio. Slope = -5.4016 Slope = 1.61578 9 10 11 12 13 14 0.70.91.11.31.51.71.9RO2 Concentration (%)73 74 75 76 77 78 79N2 Concentration (%) O2 N2 Figure 6-21. Oxygen and nitrogen concentr ations versus recirculation ratio.

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103 y = 3.27x2 9.14x + 6.50 R2 = 0.6910 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.01.21.41.61.82.0LPC Pressure RatioWFR WFR Modeled Figure 6-22. Water-to-fuel ratio versus LPC pressure ratio. y = -1.98x2 + 5.57x 2.91 R2 = 0.8400.6 0.7 0.8 0.9 1.0 1.1 1.2 1.01.11.21.31.41.51.6LPC Pressure RatioRecirculation Mass Flow Rate (lbm/s) Figure 6-23. Recirculation flow rate versus LPC pressure ratio.

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104 y = -12.5x2 + 59.6x + 47.3 R2 = 0.968 y = -56.7x2 + 217x 95.5 R2 = 0.77 y = 74.6x2 64.8x + 113 R2 = 0.81130 50 70 90 110 130 150 170 190 210 1.01.11.21.31.41.51.6LPC Pressure RatioWater Extraction Related Temperatures (F) CGC Exit CGC Inlet Dew-Point, CGC Inlet Figure 6-24. Temperatures related to wate r extraction versus LPC pressure ratio.

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105 CHAPTER 7 CONCLUSIONS AND RECOMMENDATIONS Summary Following these experimental efforts, several conclusions can be drawn about the overall success of the study, and about the theoretical cap abilities offered by the HPRTE/VARS combined cycle. Throughout this research, f our engine configurati ons were tested and progressively altered in order to arrive at an experimental apparatus best suited for operating the HPRTE/VARS, and for demonstrati ng its qualities. It was show n that the VARS hardware can be successfully integrated with pressurized a nd semi-closed engine ducting, and driven by the hot recirculating gases, resulting in a synerg istic combination of cycles. The successful operation of the HPRTE was also demonstrated, as the HPRTE was pressurized and semi-closed to various degrees. Despite operational limitatio ns presented by certain components, data were acquired and adjusted to successfully prove the following theoretically proposed concepts: The HPRTE power output can be increased /decrease by divert ing exhaust flow through/around the LPT, varying the pressure ratio across the turbocharger, and the total mass flow rate through the engine. The HPRTE efficiency can remain approximately constant for a variety of power levels, since the dimensionless operating point of the engine changes very little, or not at all during power transitions. The emission of NOX from the HPRTE is reduced as R is increased. The mass flow rates of air into and exhaust gas out of the HPRTE are reduced as R is increased. Significant quantities of water can be condens ed and extracted from the recirculating gases. The performance of the HPRTE/VARS can be hi ndered by certain of its components. The pressure drop in the recirculation line wa s the primary detraction from the power-making potential of the engine. On this engine rig, the pressure difference between the HPT exit and

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106 HPC inlet significantly reduced the HPT pressure ra tio, raising the HPT exit temperature near to its limit. When this occurred, there was little or no margin remaining in HPT temperatures to apply a load using the dynamometer. Gas leak s in the ducting and engi ne casing were also suspected to have an adverse effect on net power though the extent to which they detracted from the total flow rate remains uncertain. While an ideal engine would operate at a constant speed throughout the pressurized regime, this engine, equipped with a mechanical governor, was subject to a reduction in speed as the engine was pressurized. The speed-droop was small, but discernable. A small decrease in corrected flow rate was observed at the engine speed lessened, as well as a drop in both the HPC and HPT pressure ratios. With the drop in pr essure ratios followed a drop in both temperature ratios, and in turn the temperature differences (proportional to net power) across the HPC and HPT. Despite the departure from its design speed, the HPRTE/VARS suffered only a small decrease in thermal efficiency, barley evident in the data scatter. The VARS described in Chapter 3 was designed specifically for this application, but was intended for a slightly different ducting configur ation, feeding cooler gases to the HGC. In actuality, the HGC inlet gases were approximate ly 300-400 F (167-222 K) hotter than designed for, increasing the HGC effectiveness. This effectively shifted the optimal VARS operating point to much lower HGC and CGC flow rates, and caused the VARS to be overwhelmed at higher heat loads. This effect prohibited significant water extracti on rates from being realized at higher LPC pressure ratios, because the CGC temper atures rose much faster than the dew-point temperature. Substantial water extraction wa s however accomplished at lower LPC pressure ratios where R was highest and th e VARS performed respectably.

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107 Recommendations Through the course of these HPRTE/VARS e xperiments, much was learned about the positive attributes of the combined cycle, and ab out its limitations. The range of operations and performance of this combined cycle were found to be limited by several de tails associated with its hardware. Difficulties and uncertainties were also encountered while completing the data reduction and interpretation due to inadequacies in the instrumentation. These experiences have served to generate several recommendations that would improve the quality of future HPRTE/VARS experimental endeavors. Significant improvements can result from utilizing a VARS having a wider range of operations. This is believed as key since the HP RTE itself is capable of operating with a variety of flow rates, temperatures, and pressures in th e recirculation flow path An ideal VARS would posses more capacity and a more sophisticated cont rol scheme to maintain a constant HPC inlet temperature, and perhaps excess refrigeration cap acity. This would most strongly impact the water extraction rate from the HPRTE. While the HPC inlet (CGC exit) temperature remains constant, the dew-point te mperature would increase with LPC pressure ratio, increasing the water extraction rate. The production of excess refrigeration from the VARS would also be a favorable quality, and would allow for a new dimension of opera tions to be investigated. Specifically, this would permit experiments where excess refrigerati on can be traded for engine efficiency and water extraction, or vice versa. The potential of an HPRTE to make power coul d be greatly increased if the pressure drops in the recirculation line are kept minimal. The VARS heat exchangers used in these experiments were relatively compact, and it would be tolerable in research-oriented applications if they were somewhat larger. This would serve to reduce the back-pressure on the HPT and restore some margin in the HPT exit temperature for engine loading. Another point worth noting is that

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108 engines with higher pressure ra tios should suffer less from a give n differential pressure across the recirculation line. These experiments saw engi ne recirculation different ial pressures of up to 4 psi (28 kPa), which resulted in a 20% reduc tion in the HPT pressure ratio. The same differential pressure with a higher HPT pressure ratio should be a less sign ificant detraction from the HPT temperature difference. The range of HPRTE operation could also be markedly enhanced by employing a variable geometry turbocharger (VGT). VGTs have the ability to vary the AR ratio on their turbine housings, allowing for a variety of LPT flow rates for a given LPT pressure ratio. This would all but eliminate the challenge of selecting one turboc harger with a constant AR ratio to match the variety of recirculation ratios a nd LPC pressure ratios seen by the HPRTE. A VGT would also most likely eliminate the need for a recirculatio n valve. The recirculation valve has functioned primarily to restrict recirculat ion flow after the wastegate (VEXH) is initially closed, diverting more flow through the LPT. When operating in this way, the recircul ation pressure drop is higher than necessary. A VGT would functi on with a larger AR ratio during initial pressurization, and may further do away with the need for a wasteg ate. It may be preferable, however, to retain the service of a wastegate as it would allow multiple R-versus-pressure-ratio lines on which to operate. Future HPRTE configurations may also benefit from additional water extraction points directly upstream of the HPC in let. Condensate was observed on several occasions leaking from flanges in this location, confirming the presence of liquids in the ducting near the HPC inlet. The gas entering the HPC can be saturated with wa ter vapor (ideally it always is) and has even been observed as misty with tiny, suspended wa ter droplets. While the small droplets entering the HPC can be a performance enhancement, they can also present a problem if they coalesce on

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109 the ducting walls, or become too large for the HPC rotor being used. This can result unnecessary erosion of the compressor wheel. The last recommendation following this work relates to instrumentation, and would be found most useful in app lications where the water extraction pot ential is exploi tedthe inclusion of a duct-mounted humidity sensor. This inst rument would significantly increase the ease of data reduction, eliminating many of the approximati ons and assumptions discussed in Chapter 5. A more reliable knowledge of this quantity would result in a much more accurate dew-point temperature, heat exchanger effectiveness, and overall quantificati on of water extraction performance.

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110 APPENDIX A OPERATING PROCEDURES FOR HPRTE/VARS Starting Procedure: Prior to Start: 1 Turn on the VCR Recorder, turn VHS Camera to CAMERA 2. Do a communications check. 3. Plug in ARU. Place the ARU in standby by ro tating spring return switch clockwise. The YELLOW light should come on. 4. Turn on the condensate removal pumps. Engine Start: 1. Turn on the DC circuit breakers local to the skid. Red indicato r light should come on locally and in the Control Room 2. Move the Speed trim Valve to half way between open and shut. 3. Check open the High Pressure Fuel Manual Is olation Valve. Valve handle cross-line with the fuel pressure line. 4. Move S1 to ON. 5. Check Oil Pump #1 to Bypass. 6. Check Fuel Pump to Bypass (S9). 7. Check Oil Pump #2 to Normal. 8. Turn Oil Pump #2 to On. Confirm pressure light on before continuing. 9. Check fuel pump Off. 10. Check Air Pressure light on. 11. Move Ignition switch to ON. This enab les the ignition system and the starter 12. Check the 12VDC starter Off. 13. Check Spray Cooler switch Off.

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111 14. Check the Combustor Air Valve Off. 15. Check the solenoid Fuel Valve Off. 16. Press the RUN button. The lighted button a nd the control room i ndicator light should come on. 17. Press and Hold the START button. Check Fuel Pressure light on. The fuel valve solenoid energized light should be Off. 18. At 300 PRM (or 40 PSIG fuel pressure), turn the Solenoid Fuel Valv e on. Verify ignition and increase in engine speed. 19. Release the start button when green light (starter solenoid en ergized) goes out. 20. Engine speed should increase to self-sus taining, about 2700 dyno RPM. Wait until fuel has burned off in the sight gla ss, if necessary. Slowly move the Speed Trim valve to full OPEN, rotated down. Engine speed should in crease to approximately 3000 RPM as read on the Dyno. 21. Turn the DC exciter and Local switch to Off. Verify local oil pressure. Recirculation, ARU, Dyno, a nd Turbocharged Operation: 1. Open ARU V14. ARU V13 can be opened 1 turns to hasten warm up. 2. Slowly open the Recirculation Isol ation Valve to until recirc DP=1.6. 3. Partially close the Main Air Inlet Valve. MAI DP=0.5 in.-H2O oil max. Recirc DP=6.5. Recirc ratio=0.6. 4. Close the Dyno outlet valve unt il the flow is about 5 GPM a nd pressure is 28 to 30 PSIG. 5. ARU System pressure will rapidly build up. When the ammonia receiver pressure reaches 210 psig, open the condenser cooling water va lve to maintain pressure around 235 psig, but below 250 psig. Run data shows about 4 gpm is appropriate at full dyno load (no boost). Discharge weak solution if required. 6. Adjust solution flow to 3.2 gpm by throttling V13 if required. 7. Slowly adjust the Boost c ontrol pressure to 8.0 PSIG. Ve rify Boost Control Valve stem position is 28% open. To avoid burner instabil ity use the Recirculation Control Valve to keep R less than 1.0. Flame transitio n should occur near this setting.

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112 Recirculation Setting and Steady-State Operations: 1 Adjust HRVG Heat Load to 23TR using R ecirc and Boost control valves. During this time ARU should be adjusted to limit T-HPCI below 90F. Record data as necessary. Purge non-condensables and remove solution as necessary. PR LPT may be 1.0. 2. Increase HRVG heat load within stable limits. Record data Column 3 and 4. Recirculation control and Boost control maybe used to achieve PR LPT up to 1.8. Shut Down: 1. Remove Dyno load slowly from engine. 2. Open Boost Control valve to 7.5 PSIG. Re strict Recirc control to DP=16 (about 45 degrees). Open MAI. 3. Reduce engine speed to 2900 RPM using the Speed Trim Valve. 4. Close the Manual Fuel Shut Off Valve. 5. Press the STOP button. 6. Move the Fuel Valve switch to Off. 7. Turn Off VCR. Post Shut Down: 1. Limit soak-back to under 3800F. Motor the engine for 30 seconds as necessary. 2. Turn ARU off. Close ARU valves V14 and V2. 3. Collect Data sheets. 4. Review Tape immediately. Synchronize data sheets and data stream with unusual events. 5. Critique and review the test run. Check for signs of fluid leaks. Reconfigure the engine for a quick restart.

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113 APPENDIX B SETUP PROCEDURES FOR HPRTE/VARS Support Systems: Site Chill Water: 1. Slowly open the source and re turn overhead isolation valves. Cooler Chill Water: 1. Connect the process water hoses to the cooler. 2. Supply water to the coolers by opening all isolation valves. 3. Verify main cooler flow by listening for flow noise. Record initial flows: 4. Record minor leaks for later resolution. ARU Chill Water: 1. Connect the process water hoses to the ARU. 2. Supply water to the cooler by opening all isolation valves. 3. Verify main cooler flow by listening for flow noise. 4. Record minor leaks for later resolution Boost Control Valve and Waste Gate Control Air: 1. Set supply air regulator by the South door to 40 psig. 2. Ensure the Supply, Boost Control an d Waste Gate Bleed toggles are closed. 3. Ensure the Waste Gate Isolation toggles are open. 4. Verify the Supply Air Pressure at the control panel is between 25 and 30 psig by adjusting the South door regulator in 1. above. 5. Open the Fisher Control Regulator until th e Boost Control Air gage reads 15 psig. This opens the 6 Boost Cont rol valve completely. 6. Visually verify the position of the Boost Control Valve is full open by observation of the valve stem indicator.

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114 7. Verify the turbocharger Waste Gate Regulator is set to 20 psig. 8. Verify operation of the Turboc harger Waste Gate with an as sistant listening for actuation, from full open to full closed, by closing the Waste Gate Isolation Toggle and opening the Waste Gate Bleed Toggle. 9. Reset the toggles for Waste Gate position Full open : Isolation Toggle Open/Bleed Toggle closed. 12 Volt Battery Check 1 Unplug the battery charge r and store the charger. 2. Throw the isolation switch. 3. Verify each start battery has a cold reading of 13.2 volts mi nimum. This ensures that the batteries are fully char ged. Record Voltage: 4. Verify each ignition battery has a cold read ing of 12.6 volts minimum. This ensures that the batteries are sufficiently charged. Record Voltage: Fuel Supply 1. Verify Gravity and speed trim lines to engi ne skid. Open the ROVER side of the gravity system. Place the speed trim valve in the ope n position, rotated down. Shutting this valve will trim engine speed and eventually shut the engine down. 2. Purge the fuel Pump accumulator to charge with air and rig for remote valve operation. 3. Using the battery charger AC supply, plug in the fuel transfer pu mp and verify High and Low speed operation. 4. Check fuel level by sight glass and control panel gage, minimum full. Fill as necessary. 5. Fill the seven gallon tank under the fuel cabinet, if required. 6. Check the entire fuel cabinet and hoses for leaks. 7. Have a full 5 gallon fuel can standing by in the fuel closet, if required. 8. Dry and position the skid drip pa n and the fuel drain drip pan. Oil Levels

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115 1. Verify Rover engine oil level. Refill with single viscosity 10W oil as necessary up to half way between the high and low markings. 2. Safety wire the dipstick. 3. Verify Turbocharger oil level. Fill with 10W -30 oil as necessary as not to flood the turbo oil scavenge port. Visually ch eck the condition of the oil. Dyno Setup and Oil Cooler Process Water 1. Crack open the Turbocharger Lube Oil Cooler water supply tap. 2. Route system drain lines outsi de under the main overhead door. 3. Fully open the water supply to the Froude Dyno. Purge Dyno. 4. Verify the in-line Rota meter is reading 6.9 GPM. Adjust Dyno in let and outlet valve to full open. 5. Verify the water brake is fully unloaded, th at is, the geared handle is full to the CCW position. 6. Verify the gear lock remains disengaged. 7. Visually confirm gland leakage. 8. Confirm discharge flow from four drain line s. They are; 1.)The Rover oil cooler, 2.) The dyno, 3.) The dyno drip pan and 4.) Th e Turbocharger Lube Oil Cooler. Water Recovery 1. Attach the Water Recovery Reservoir to the load cell. 2. Supply AC power to the drain pumps and turn them on. Verify rotation

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116 Inlet, Recirculation a nd Exhaust Start-up Check 1. Check the exhaust system to ensure all pe netrations are covered and joints are tight. 2. Verify that the Boost Control Valve is fully open. 3. Verify that the Rover Inle t Isolation Valve is fully open 4. Verify the Waste Gate Valve is fully open. 5. Verify the Rover Recircul ation Valve is fully shut. Engine room Preparation Room Ventilation: 1. Open the main bay door about five feet (to marked line). 2. Turn on the lab ventilation fan and the compressor room fan. 3. Turn the air conditioning thermostat bypass, at the north wall, to on. Lab Over-watch: 1. Check the lab area for debris that could be i ngested into the engine or present a tripping hazard. 2. Attach the Safety Chain at the hall outside the Lab. 3. Move fire extinguishers to areas in the lab where they are readily accessible. 4. Visitor Policy: all visitors should be checked in, briefe d, and supplied with safety equipment before the run set-up begins. Optim ally, all visitors should be supplied with Listen-Only communication head gear. No la te or unannounced visitors are allowed. 5. Put the scatter shields in place fo r turbocharged runs, if required. Gas Analysis and Data Acquisition 1. Refer to the gas analysis set up pro cedure, separate from this document. 2. Verify that all thermoc ouples are reading properly by bot h the analog and digital data acquisition systems. Confirm conformance to the instrumentation map. This should be completed a day in advance of the run.

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117 3. Verify that all pressure taps are reading properly by bot h the analog and digital data acquisition systems. Confirm conformance to the instrumentation map. This should be completed a day in advance of the run. 4. Verify the analog pressure, temperature, and manometer reading legends are clearly displayed on the panel near the instruments. 5. Record the initial reading from the Wa ter Recovery load cell on the Data Sheet. ARU Setup 1. Check if the solution receiver is at least one quarter full. If it is lower than 1/4, see troubleshooting guide. 2. Check all the valve positi ons according to valve tag list. 3. Check closed V13 wide (pump bypass). 4. Record the initial So lution Receiver tank level. 5. Record the initial NH3 Receiver Tank level. 6. Set V14 shut (pump discharge). 7. V15 at 3-turn open, Set by Supplier (Mattingly) (col feed). 8. Adjust water valves V26 and V27 to get 2 gpm in condenser and 10 gpm in absorber. Record initial flows: 9. Turn power switch on. Panel lights should go from Red to Yellow until HRVG inlet air becomes hot. Then the Green light will come on starting the solution pump and opening the solenoid valves. Video and Audio Recording 1. Set up the VCR in the lab. Run the cables to the control room to the video monitor. 2. Set up the microphone to record the communi cations loop. Hook this into the VCR sound input. 3. Insert a new VCR tape for the days activity Ensure tape is recording on E.P. (extended play). 4. Synchronize TIME and DATE. 5. Complete a system check to be sure that all the monitoring and recording systems are working correctly. This system is used to ve rify the data set switch points and aids in improving each subsequent runs through lessons learned. Personnel Safety Equipment and Communications

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118 1. All personnel should wear appropriat e clothing for an environment where high temperature piping, heavy equipment and high speed rotating equipment exist, i.e., long sleeve shirt, long pants, closed -toe and heel shoes, and no loose fitting items or jewelry. TURN OFF CELL PHONES. 2. Check all the communication gear. Check batteries. All un its should be on the same channel and in Push-To-Talk (PTT) mode. All units should be on TX, not INT. Use channel A. 3. All personnel and visitors should have hearing protection, either communi cation sets or ear muffs. 4. All personnel and visitors should have eye protection.

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119 APPENDIX C ADJUSTED EXPERIMENTAL DATA Provided below is the reduced data for all of the successful Build 4 engine runs. Each engine run is identified as B4 -X, where X represents the run number. Each run had between one and five separate trials. The reduced para meters shown were calculated after the data adjustments were made, according to the discussion in Chapter 5. Table C-1. Data from runs B4-1 and B4-2. Run B4-1 B4-2 Trial 1 2 1 2 3 Main Air Inlet Ambient Temperature (F) 69.3 70.7 84.9 86.5 87.8 Ambient Pressure (psia) 14.7 14.7 14.7 14.7 14.7 Ambient Humidity Ratio 0.01890.01890.01710.0174 0.0174 Mass Flow Rate (lbm/s) 0.00 0.00 0.00 0.00 0.00 Low Pressure Compressor Inlet Temperature (F) 96.0 96.0 95.6 97.3 98.7 Gamma 1.40 1.40 1.40 1.40 1.40 Shaft Power (hp) 2.8 4.3 8.9 7.9 13.0 Isentropic Efficiency (%) 54.9 54.8 66.6 71.0 73.3 Pressure Ratio 1.06 1.08 1.18 1.18 1.28 Flow Rate (lbm/s) 0.52 0.54 0.64 0.60 0.68 Corrected Flow Rate (lbm/s) 0.54 0.56 0.66 0.62 0.70 High Pressure Compressor Inlet Temperature (F) 76.9 77.7 76.6 79.7 89.2 Inlet Pressure (psia) 15.2 15.6 17.1 17.1 18.5 Adiabatic Exit Temperature (F) 330.4 330.3 327.3 329.9 339.2 Gamma 1.38 1.38 1.38 1.38 1.38 Shaft Power (hp) 115.4 119.2 131.0 130.9 139.4 Isentropic Efficiency (%) 70.8 71.3 71.5 71.7 70.2 Pressure Ratio 2.85 2.86 2.82 2.81 2.72 Total Gas Flow Rate (lbm/s) 1.31 1.35 1.49 1.49 1.59 Corrected Flow Rate (lbm/s) 1.29 1.30 1.31 1.31 1.30 Combustor Fuel Flow Rate (lbm/s) 0.018 0.019 0.020 0.020 0.020 Phi 0.50 0.49 0.44 0.47 0.43 Combustion Efficiency 91.7 91.7 96.6 96.6 97.3 Combustor Exit Temperature (F) 1472 1463 1485 1474 1480 High Pressure Turbine Inlet Temperature (F) 1429 1430 1414 1400 1419 Inlet Pressure (psia) 43.3 44.4 48.2 48.0 50.3 Gamma 1.31 1.31 1.32 1.32 1.32

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120 Run B4-1 B4-2 Trial 1 2 1 2 3 Shaft Power (hp) -180.0 -186.4 -204.1 -202.1 -211.4 Isentropic Efficiency (%) 85.0 85.0 85.0 85.0 85.0 Expansion Ratio 2.85 2.86 2.82 2.81 2.72 Corrected Flow Rate (lbm/s) 0.82 0.83 0.84 0.85 0.86 Recuperator Hot Side dT (F) -176.6 -165.9 -120.4 -122.5 -136.1 Hot Side Heat Rate (Btu/s) -78.5 -82.6 -88.5 -85.7 -89.9 Cold Side dT (F) 218.3 223.2 217.2 210.0 207.4 Cold Side Heat Rate (Btu/s) 72.6 76.7 82.5 79.9 84.2 Heat to Ambient (Btu/s) -5.9 -5.9 -5.9 -5.8 -5.7 Effectiveness 0.32 0.32 0.32 0.31 0.32 Low Pressure Turbine Inlet Temperature (F) 899 910 938 926 937 Inlet Pressure (psia) 17.5 17.9 19.4 19.1 20.5 Gamma 1.33 1.33 1.34 1.34 1.34 Shaft Power (hp) -2.8 -4.3 -8.9 -7.9 -13.0 Isentropic Efficiency (%) 35.8 45.5 53.2 52.1 58.7 Expansion Ratio 1.19 1.21 1.32 1.30 1.39 Flow Rate (lbm/s) 0.36 0.39 0.47 0.45 0.52 Corrected Flow Rate (lbm/s) 0.49 0.52 0.58 0.57 0.61 Hot Gas Cooler Gas Side dT (F) -427.7 -431.6 -431.8 -412.6 -410.4 Gas Side Heat Rate (Btu/s) -91.4 -94.1 -98.8 -98.0 -99.9 Warm Gas Cooler Gas Side dT (F) -393.4 -389.9 -405.6 -398.5 -405.3 Water Side dT (F) 11.0 12.1 12.5 13.3 14.1 Gas Side Heat Rate (Btu/s) -86.8 -85.6 -91.9 -93.6 -96.4 Water Side Heat Rate (Btu/s) 59.1 61.5 65.8 69.1 73.2 Heat Loss to Ambient (Btu/s) -27.7 -24.1 -26.1 -24.4 -23.2 Effectiveness 0.73 0.71 0.70 0.69 0.68 Cold Gas Cooler Gas Side dT (F) -49.9 -52.0 -58.6 -61.2 -58.3 Gas Side Heat Rate (Btu/s) -23.1 -22.2 -25.1 -25.9 -25.2 Overall Performance R 1.53 1.51 1.35 1.48 1.35 Power (hp) 64.5 67.2 73.1 71.2 72.0 Thermal Efficiency (%) 13.90 13.97 14.40 14.05 13.64 Average Water Extraction Rate (gph)5.81 4.00 2.88 2.69 1.86 Average Fuel Flow Rate (gph) 9.21 9.54 10.08 10.06 10.47

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121 Table C-2. Data from runs B4-3 and B4-5. Run B4-3 B4-5 Trial 1 2 3 1 2 Main Air Inlet Ambient Temperature (F) 85.2 86.5 87.2 86.0 87.0 Ambient Pressure (psia) 14.8 14.8 14.8 14.7 14.7 Ambient Humidity Ratio 0.01860.01860.01860.0169 0.0169 Mass Flow Rate (lbm/s) 0.00 0.00 0.00 0.00 0.00 Low Pressure Compressor Inlet Temperature (F) 95.0 95.0 95.0 95.5 97.8 Gamma 1.39 1.39 1.40 1.40 1.40 Shaft Power (hp) 16.7 18.3 2.7 15.8 17.1 Isentropic Efficiency (%) 72.4 71.4 39.4 77.7 78.8 Pressure Ratio 1.34 1.36 1.04 1.37 1.39 Flow Rate (lbm/s) 0.73 0.74 0.50 0.69 0.70 Corrected Flow Rate (lbm/s) 0.75 0.77 0.51 0.71 0.73 High Pressure Compressor Inlet Temperature (F) 93.6 102.7 71.5 97.2 108.3 Inlet Pressure (psia) 19.3 19.7 15.0 19.8 20.1 Adiabatic Exit Temperature (F) 342.4 350.6 326.3 345.9 355.9 Gamma 1.38 1.38 1.38 1.38 1.38 Shaft Power (hp) 145.0 144.4 116.9 149.1 147.9 Isentropic Efficiency (%) 70.7 70.6 70.9 69.5 69.4 Pressure Ratio 2.70 2.65 2.87 2.65 2.60 Total Gas Flow Rate (lbm/s) 1.66 1.66 1.31 1.71 1.70 Corrected Flow Rate (lbm/s) 1.30 1.29 1.30 1.31 1.30 Combustor Fuel Flow Rate (lbm/s) 0.021 0.021 0.018 0.021 0.021 Phi 0.41 0.40 0.52 0.44 0.43 Combustion Efficiency 97.6 98.0 96.1 96.2 97.3 Combustor Exit Temperature (F) 1477 1495 1514 1435 1458 High Pressure Turbine Inlet Temperature (F) 1423 1437 1406 1398 1416 Inlet Pressure (psia) 52.3 52.2 43.1 52.4 52.3 Gamma 1.32 1.32 1.32 1.32 1.32 Shaft Power (hp) -220.4 -218.1 -181.7 -219.7 -216.8 Isentropic Efficiency (%) 85.0 85.0 85.0 85.0 85.0 Expansion Ratio 2.70 2.65 2.87 2.65 2.60 Corrected Flow Rate (lbm/s) 0.86 0.86 0.84 0.87 0.87 Recuperator Hot Side dT (F) -128.5 -132.2 -127.4 -140.3 -143.8 Hot Side Heat Rate (Btu/s) -91.5 -95.2 -75.9 -91.6 -93.7 Cold Side dT (F) 201.5 210.2 209.4 196.4 202.1 Cold Side Heat Rate (Btu/s) 85.5 89.2 69.9 85.7 87.9 Heat to Ambient (Btu/s) -6.0 -6.0 -6.0 -5.9 -5.8 Effectiveness 0.32 0.32 0.30 0.32 0.32

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122 Run B4-3 B4-5 Trial 1 2 3 1 2 Low Pressure Turbine Inlet Temperature (F) 950 964 921 923 941 Inlet Pressure (psia) 21.5 21.8 17.0 21.2 21.5 Gamma 1.34 1.34 1.34 1.34 1.34 Shaft Power (hp) -16.7 -18.3 -2.7 -15.8 -17.1 Isentropic Efficiency (%) 61.4 63.1 21.6 61.5 61.7 Expansion Ratio 1.45 1.47 1.16 1.44 1.46 Flow Rate (lbm/s) 0.56 0.57 0.32 0.55 0.57 Corrected Flow Rate (lbm/s) 0.63 0.64 0.46 0.63 0.63 Hot Gas Cooler Gas Side dT (F) -418.5 -420.4 -409.8 -383.6 -389.9 Gas Side Heat Rate (Btu/s) -104.0 -102.6 -90.5 -104.4 -103.5 Warm Gas Cooler Gas Side dT (F) -412.2 -418.0 -399.8 -424.7 -435.6 Water Side dT (F) 14.8 15.0 11.7 16.2 16.3 Gas Side Heat Rate (Btu/s) -100.1 -99.2 -91.8 -112.4 -112.1 Water Side Heat Rate (Btu/s) 74.0 75.1 65.2 79.1 81.8 Heat Loss to Ambient (Btu/s) -26.0 -24.1 -26.6 -33.3 -30.4 Effectiveness 0.69 0.68 0.72 0.70 0.70 Cold Gas Cooler Gas Side dT (F) -51.0 -46.2 -61.4 -46.1 -38.7 Gas Side Heat Rate (Btu/s) -23.4 -21.0 -27.6 -21.5 -18.1 Overall Performance R 1.28 1.23 1.66 1.49 1.42 Power (hp) 75.4 73.7 64.8 70.5 68.9 Thermal Efficiency (%) 13.85 13.59 13.81 12.96 12.76 Average Water Extraction Rate (gph)1.97 1.64 6.27 1.67 1.49 Average Fuel Flow Rate (gph) 10.80 10.76 9.31 10.80 10.71

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123 Table C-3. Data from runs B4-7 and B4-8 Run B4-7 B4-8 Trial 1 2 3 4 1 Main Air Inlet Ambient Temperature (F) 85.1 86.0 87.7 88.1 84.2 Ambient Pressure (psia) 14.8 14.8 14.8 14.8 14.7 Ambient Humidity Ratio 0.01860.01860.01900.0200 0.0189 Mass Flow Rate (lbm/s) 0.00 0.00 0.00 0.00 0.00 Low Pressure Compressor Inlet Temperature (F) 90.3 92.5 93.3 94.1 90.1 Gamma 1.40 1.40 1.40 1.39 1.39 Shaft Power (hp) 18.1 13.3 14.9 15.6 27.7 Isentropic Efficiency (%) 69.8 72.5 73.3 73.3 69.6 Pressure Ratio 1.36 1.30 1.33 1.34 1.48 Flow Rate (lbm/s) 0.73 0.65 0.67 0.68 0.86 Corrected Flow Rate (lbm/s) 0.75 0.66 0.69 0.70 0.88 High Pressure Compressor Inlet Temperature (F) 88.7 92.2 92.3 93.3 89.5 Inlet Pressure (psia) 19.7 18.9 19.3 19.5 21.4 Adiabatic Exit Temperature (F) 337.5 342.3 339.4 340.1 331.7 Gamma 1.38 1.38 1.38 1.38 1.39 Shaft Power (hp) 151.2 146.0 147.1 148.1 157.5 Isentropic Efficiency (%) 70.7 70.3 71.0 70.8 70.5 Pressure Ratio 2.73 2.71 2.71 2.69 2.65 Total Gas Flow Rate (lbm/s) 1.73 1.66 1.69 1.70 1.85 Corrected Flow Rate (lbm/s) 1.33 1.33 1.33 1.33 1.31 Combustor Fuel Flow Rate (lbm/s) 0.021 0.020 0.023 0.024 0.023 Phi 0.42 0.46 0.49 0.50 0.38 Combustion Efficiency 96.1 94.9 96.8 97.2 99.6 Combustor Exit Temperature (F) 1427 1417 1521 1545 1473 High Pressure Turbine Inlet Temperature (F) 1387 1373 1414 1415 1418 Inlet Pressure (psia) 53.6 51.2 52.2 52.5 56.6 Gamma 1.32 1.32 1.32 1.32 1.32 Shaft Power (hp) -226.8 -215.2 -223.4 -224.2 -240.9 Isentropic Efficiency (%) 85.0 85.0 85.0 85.0 85.0 Expansion Ratio 2.73 2.71 2.71 2.69 2.65 Corrected Flow Rate (lbm/s) 0.86 0.86 0.89 0.90 0.88 Recuperator Hot Side dT (F) -98.3 -98.5 -108.3 -105.3 -109.4 Hot Side Heat Rate (Btu/s) -96.1 -90.6 -97.1 -98.0 -110.9 Cold Side dT (F) 203.3 198.9 210.3 210.7 221.0 Cold Side Heat Rate (Btu/s) 89.8 84.4 91.0 92.0 104.6 Heat to Ambient (Btu/s) -6.3 -6.2 -6.1 -6.0 -6.3 Effectiveness 0.33 0.33 0.30 0.29 0.33

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124 Run B4-7 B4-8 Trial 1 2 3 4 1 Low Pressure Turbine Inlet Temperature (F) 947 937 964 970 971 Inlet Pressure (psia) 21.6 20.3 20.8 21.0 23.9 Gamma 1.34 1.34 1.34 1.34 1.34 Shaft Power (hp) -18.1 -13.3 -14.9 -15.6 -27.7 Isentropic Efficiency (%) 64.4 62.7 63.4 63.9 67.6 Expansion Ratio 1.46 1.38 1.41 1.42 1.62 Flow Rate (lbm/s) 0.57 0.51 0.52 0.53 0.65 Corrected Flow Rate (lbm/s) 0.64 0.60 0.61 0.62 0.67 Hot Gas Cooler Gas Side dT (F) -391.9 -390.7 -394.5 -393.2 -413.8 Gas Side Heat Rate (Btu/s) -104.1 -105.3 -107.5 -107.6 -109.5 Warm Gas Cooler Gas Side dT (F) -412.5 -411.0 -437.4 -441.0 -422.0 Water Side dT (F) 16.1 15.8 15.8 16.2 16.3 Gas Side Heat Rate (Btu/s) -105.1 -106.1 -114.1 -115.5 -107.1 Water Side Heat Rate (Btu/s) 78.6 78.6 87.2 92.3 79.2 Heat Loss to Ambient (Btu/s) -26.4 -27.5 -26.9 -23.2 -27.9 Effectiveness 0.66 0.67 0.69 0.69 0.67 Cold Gas Cooler Gas Side dT (F) -71.0 -58.9 -59.5 -60.6 -76.8 Gas Side Heat Rate (Btu/s) -32.0 -25.5 -26.8 -27.5 -37.7 Overall Performance R 1.37 1.57 1.52 1.51 1.16 Power (hp) 75.6 69.1 76.3 76.1 83.4 Thermal Efficiency (%) 13.72 13.04 12.77 12.39 14.13 Average Water Extraction Rate (gph)0.81 0.66 0.96 0.94 1.24 Average Fuel Flow Rate (gph) 10.94 10.52 11.86 12.20 11.71

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125 Table C-4. Data from runs B4-10 and B4-11. Run B4-10 B4-11 Trial 1 2 1 2 Main Air Inlet Ambient Temperature (F) 86.1 86.6 83.8 84.2 Ambient Pressure (psia) 14.8 14.8 14.8 14.8 Ambient Humidity Ratio 0.01860.01830.01690.0174 Mass Flow Rate (lbm/s) 0.00 0.00 0.00 0.00 Low Pressure Compressor Inlet Temperature (F) 93.0 93.9 91.5 92.5 Gamma 1.40 1.40 1.40 1.40 Shaft Power (hp) 14.2 14.3 16.8 17.8 Isentropic Efficiency (%) 67.3 67.9 67.7 68.0 Pressure Ratio 1.29 1.30 1.34 1.36 Flow Rate (lbm/s) 0.65 0.65 0.69 0.70 Corrected Flow Rate (lbm/s) 0.67 0.67 0.71 0.72 High Pressure Compressor Inlet Temperature (F) 88.2 89.4 88.8 91.2 Inlet Pressure (psia) 19.1 19.2 18.9 20.0 Adiabatic Exit Temperature (F) 339.3 340.5 338.7 340.7 Gamma 1.38 1.38 1.38 1.38 Shaft Power (hp) 153.3 152.3 148.5 150.1 Isentropic Efficiency (%) 70.6 70.3 71.2 71.5 Pressure Ratio 2.75 2.73 2.75 2.75 Total Gas Flow Rate (lbm/s) 1.73 1.72 1.69 1.71 Corrected Flow Rate (lbm/s) 1.37 1.36 1.29 1.29 Combustor Fuel Flow Rate (lbm/s) 0.021 0.022 0.022 0.023 Phi 0.47 0.48 0.47 0.47 Combustion Efficiency 94.4 95.2 97.4 97.1 Combustor Exit Temperature (F) 1399 1435 1500 1496 High Pressure Turbine Inlet Temperature (F) 1354 1363 1409 1403 Inlet Pressure (psia) 52.5 52.4 54.4 55.0 Gamma 1.32 1.32 1.32 1.32 Shaft Power (hp) -225.0 -223.4 -225.9 -227.5 Isentropic Efficiency (%) 85.0 85.0 85.0 85.0 Expansion Ratio 2.75 2.73 2.75 2.75 Corrected Flow Rate (lbm/s) 0.88 0.88 0.85 0.85 Recuperator Hot Side dT (F) -92.3 -95.4 -99.4 -99.0 Hot Side Heat Rate (Btu/s) -93.6 -94.2 -97.3 -98.3 Cold Side dT (F) 197.3 199.9 210.8 210.9 Cold Side Heat Rate (Btu/s) 87.5 88.2 91.1 92.2 Heat to Ambient (Btu/s) -6.1 -6.1 -6.2 -6.2 Effectiveness 0.32 0.31 0.30 0.30

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126 Run B4-10 B4-11 Trial 1 2 1 2 Low Pressure Turbine Inlet Temperature (F) 923 930 963 959 Inlet Pressure (psia) 20.7 20.8 21.3 21.5 Gamma 1.34 1.34 1.34 1.34 Shaft Power (hp) -14.2 -14.3 -16.8 -17.8 Isentropic Efficiency (%) 63.8 63.7 63.3 63.8 Expansion Ratio 1.39 1.39 1.44 1.46 Flow Rate (lbm/s) 0.52 0.52 0.55 0.56 Corrected Flow Rate (lbm/s) 0.61 0.61 0.63 0.63 Hot Gas Cooler Gas Side dT (F) -370.2 -375.4 -384.3 -382.9 Gas Side Heat Rate (Btu/s) -106.6 -106.8 -103.0 -103.3 Warm Gas Cooler Gas Side dT (F) -417.2 -417.9 -442.3 -434.2 Water Side dT (F) 16.4 16.1 17.7 22.5 Gas Side Heat Rate (Btu/s) -115.8 -114.4 -113.9 -112.7 Water Side Heat Rate (Btu/s) 81.4 82.6 85.2 82.6 Heat Loss to Ambient (Btu/s) -34.3 -31.8 -28.7 -30.1 Effectiveness 0.68 0.68 0.69 0.68 Cold Gas Cooler Gas Side dT (F) -64.1 -65.8 -67.2 -69.0 Gas Side Heat Rate (Btu/s) -28.8 -29.2 -29.8 -30.9 Overall Performance R 1.66 1.63 1.46 1.44 Power (hp) 71.7 71.1 77.4 77.5 Thermal Efficiency (%) 12.96 12.56 13.32 13.11 Average Water Extraction Rate (gph)0.88 0.75 1.02 1.08 Average Fuel Flow Rate (gph) 10.98 11.23 11.53 11.73

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127 Table C-5. Data from run B4-12. Run B4-12 Trial 1 2 3 4 Main Air Inlet Ambient Temperature (F) 84.4 85.7 86.5 87.1 Ambient Pressure (psia) 14.8 14.8 14.8 14.8 Ambient Humidity Ratio 0.01800.01830.01860.0186 Mass Flow Rate (lbm/s) 0.00 0.00 0.00 0.00 Low Pressure Compressor Inlet Temperature (F) 91.3 93.7 95.0 95.8 Gamma 1.39 1.39 1.39 1.39 Shaft Power (hp) 22.2 24.1 24.9 26.3 Isentropic Efficiency (%) 67.6 68.7 69.1 69.3 Pressure Ratio 1.41 1.45 1.46 1.48 Flow Rate (lbm/s) 0.76 0.78 0.79 0.80 Corrected Flow Rate (lbm/s) 0.78 0.80 0.81 0.82 High Pressure Compressor Inlet Temperature (F) 92.6 99.5 99.8 101.9 Inlet Pressure (psia) 20.9 21.4 21.6 21.9 Adiabatic Exit Temperature (F) 342.8 350.1 349.2 351.6 Gamma 1.38 1.38 1.38 1.38 Shaft Power (hp) 157.1 157.9 158.0 157.8 Isentropic Efficiency (%) 70.3 69.5 69.7 69.6 Pressure Ratio 2.71 2.66 2.65 2.64 Total Gas Flow Rate (lbm/s) 1.78 1.79 1.80 1.79 Corrected Flow Rate (lbm/s) 1.30 1.28 1.27 1.25 Combustor Fuel Flow Rate (lbm/s) 0.023 0.023 0.023 0.023 Phi 0.43 0.42 0.41 0.41 Combustion Efficiency 95.8 96.4 96.5 96.5 Combustor Exit Temperature (F) 1448 1457 1460 1473 High Pressure Turbine Inlet Temperature (F) 1392 1398 1399 1399 Inlet Pressure (psia) 56.5 56.8 57.2 57.7 Gamma 1.32 1.32 1.32 1.32 Shaft Power (hp) -233.5 -231.1 -232.3 -230.6 Isentropic Efficiency (%) 85.0 85.0 85.0 85.0 Expansion Ratio 2.71 2.66 2.65 2.64 Corrected Flow Rate (lbm/s) 0.85 0.85 0.85 0.84 Recuperator Hot Side dT (F) -98.1 -105.1 -102.1 -97.6 Hot Side Heat Rate (Btu/s) -98.8 -97.9 -98.6 -98.6 Cold Side dT (F) 202.8 200.5 201.0 201.9 Cold Side Heat Rate (Btu/s) 92.5 91.8 92.6 92.7 Heat to Ambient (Btu/s) -6.2 -6.1 -6.0 -5.9 Effectiveness 0.32 0.32 0.32 0.31

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128 Run B4-12 Trial 1 2 3 4 Low Pressure Turbine Inlet Temperature (F) 954 957 962 968 Inlet Pressure (psia) 22.6 23.1 23.3 23.5 Gamma 1.34 1.34 1.34 1.34 Shaft Power (hp) -22.2 -24.1 -24.9 -26.3 Isentropic Efficiency (%) 65.2 65.6 65.8 66.5 Expansion Ratio 1.53 1.56 1.57 1.59 Flow Rate (lbm/s) 0.61 0.63 0.64 0.65 Corrected Flow Rate (lbm/s) 0.66 0.67 0.67 0.67 Hot Gas Cooler Gas Side dT (F) -382.2 -392.4 -389.2 -394.8 Gas Side Heat Rate (Btu/s) -104.6 -105.9 -105.2 -104.8 Warm Gas Cooler Gas Side dT (F) -408.1 -403.2 -404.4 -402.2 Water Side dT (F) 17.7 18.4 18.7 18.9 Gas Side Heat Rate (Btu/s) -108.1 -105.0 -105.5 -103.2 Water Side Heat Rate (Btu/s) 81.7 83.4 84.7 85.8 Heat Loss to Ambient (Btu/s) -26.3 -21.6 -20.8 -17.4 Effectiveness 0.65 0.64 0.64 0.64 Cold Gas Cooler Gas Side dT (F) -82.4 -78.9 -83.4 -84.3 Gas Side Heat Rate (Btu/s) -37.8 -36.1 -38.3 -38.7 Overall Performance R 1.35 1.30 1.29 1.25 Power (hp) 76.4 73.2 74.3 72.9 Thermal Efficiency (%) 13.05 12.56 12.66 12.27 Average Water Extraction Rate (gph)1.07 0.85 0.85 0.90 Average Fuel Flow Rate (gph) 11.62 11.56 11.65 11.78

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129 Table C-6. Data from run B4-14. Run B4-14 Trial 2 3 4 5 Main Air Inlet Ambient Temperature (F) 84.3 85.8 86.8 88.6 Ambient Pressure (psia) 14.7 14.7 14.7 14.7 Ambient Humidity Ratio 0.01710.01760.01830.0181 Mass Flow Rate (lbm/s) 0.12 0.11 0.00 0.00 Low Pressure Compressor Inlet Temperature (F) 89.8 92.9 94.5 97.5 Gamma 1.40 1.40 1.40 1.39 Shaft Power (hp) 5.1 5.3 12.2 14.5 Isentropic Efficiency (%) 34.8 33.2 55.9 66.2 Pressure Ratio 1.06 1.06 1.20 1.30 Flow Rate (lbm/s) 0.55 0.55 0.66 0.65 Corrected Flow Rate (lbm/s) 0.57 0.57 0.68 0.67 High Pressure Compressor Inlet Temperature (F) 72.3 74.0 81.5 91.0 Inlet Pressure (psia) 15.6 15.6 17.7 19.1 Adiabatic Exit Temperature (F) 327.4 329.2 332.7 342.9 Gamma 1.39 1.39 1.38 1.38 Shaft Power (hp) 125.7 123.7 140.1 148.0 Isentropic Efficiency (%) 72.1 72.3 72.3 70.6 Pressure Ratio 2.90 2.90 2.83 2.74 Total Gas Flow Rate (lbm/s) 1.41 1.38 1.59 1.67 Corrected Flow Rate (lbm/s) 1.34 1.32 1.35 1.32 Combustor Fuel Flow Rate (lbm/s) 0.019 0.019 0.021 0.021 Phi 0.40 0.40 0.45 0.47 Combustion Efficiency 99.7 99.9 97.6 95.1 Combustor Exit Temperature (F) 1503 1528 1473 1425 High Pressure Turbine Inlet Temperature (F) 1373 1388 1383 1367 Inlet Pressure (psia) 45.4 45.3 50.1 52.5 Gamma 1.32 1.32 1.32 1.32 Shaft Power (hp) -192.7 -191.0 -214.3 -217.8 Isentropic Efficiency (%) 85.0 85.0 85.0 85.0 Expansion Ratio 2.90 2.90 2.83 2.74 Corrected Flow Rate (lbm/s) 0.85 0.84 0.86 0.85 Recuperator Hot Side dT (F) -127.0 -112.5 -85.0 -101.5 Hot Side Heat Rate (Btu/s) -83.4 -83.4 -92.5 -89.6 Cold Side dT (F) 215.5 219.6 213.8 196.3 Cold Side Heat Rate (Btu/s) 77.0 77.3 86.5 83.9 Heat to Ambient (Btu/s) -6.4 -6.1 -6.0 -5.8 Effectiveness 0.30 0.30 0.31 0.32

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130 Run B4-14 Trial 2 3 4 5 Low Pressure Turbine Inlet Temperature (F) 889 916 947 925 Inlet Pressure (psia) 17.9 17.9 20.1 20.6 Gamma 1.34 1.34 1.34 1.34 Shaft Power (hp) -5.1 -5.3 -12.2 -14.5 Isentropic Efficiency (%) -3.2 25.7 59.8 63.4 Expansion Ratio 1.21 1.22 1.37 1.41 Flow Rate (lbm/s) 0.39 0.39 0.50 0.53 Corrected Flow Rate (lbm/s) 0.52 0.52 0.60 0.61 Hot Gas Cooler Gas Side dT (F) -411.1 -417.3 -404.0 -374.5 Gas Side Heat Rate (Btu/s) -81.0 -79.7 -99.7 -102.0 Warm Gas Cooler Gas Side dT (F) -390.1 -391.7 -420.2 -398.2 Water Side dT (F) 11.7 12.6 15.6 17.5 Gas Side Heat Rate (Btu/s) -76.6 -73.8 -101.0 -104.7 Water Side Heat Rate (Btu/s) 55.6 59.6 72.9 81.3 Heat Loss to Ambient (Btu/s) -21.0 -14.2 -28.1 -23.4 Effectiveness 0.70 0.68 0.68 0.64 Cold Gas Cooler Gas Side dT (F) -52.6 -60.5 -67.1 -80.8 Gas Side Heat Rate (Btu/s) -20.0 -21.4 -28.5 -34.4 Overall Performance R 1.11 1.08 1.39 1.58 Power (hp) 67.0 67.3 74.2 69.7 Thermal Efficiency (%) 13.86 13.88 13.85 12.86 Average Water Extraction Rate (gph)2.64 1.94 1.71 0.77 Average Fuel Flow Rate (gph) 9.59 9.62 10.64 10.76

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131 Table C-7. Data from runs B4-15 and B4-16 Run B4-15 B4-16 Trial 1 2 3 1 Main Air Inlet Ambient Temperature (F) 83.0 86.1 86.5 85.1 Ambient Pressure (psia) 14.7 14.7 14.7 14.7 Ambient Humidity Ratio 0.01660.01710.01710.0186 Mass Flow Rate (lbm/s) 0.12 0.00 0.00 0.13 Low Pressure Compressor Inlet Temperature (F) 88.3 93.9 94.8 91.3 Gamma 1.40 1.40 1.39 1.40 Shaft Power (hp) 5.2 16.5 24.0 5.0 Isentropic Efficiency (%) 30.7 61.6 64.6 36.7 Pressure Ratio 1.06 1.29 1.41 1.06 Flow Rate (lbm/s) 0.54 0.70 0.79 0.55 Corrected Flow Rate (lbm/s) 0.56 0.72 0.81 0.56 High Pressure Compressor Inlet Temperature (F) 70.9 85.9 91.0 72.6 Inlet Pressure (psia) 15.5 19.0 20.7 15.6 Adiabatic Exit Temperature (F) 326.7 337.3 340.7 327.6 Gamma 1.39 1.38 1.38 1.39 Shaft Power (hp) 124.4 146.9 158.4 126.7 Isentropic Efficiency (%) 71.9 70.7 70.5 71.8 Pressure Ratio 2.91 2.76 2.71 2.89 Total Gas Flow Rate (lbm/s) 1.39 1.66 1.80 1.42 Corrected Flow Rate (lbm/s) 1.33 1.32 1.32 1.35 Combustor Fuel Flow Rate (lbm/s) 0.019 0.021 0.022 0.019 Phi 0.41 0.43 0.41 0.41 Combustion Efficiency 99.6 96.3 96.5 99.6 Combustor Exit Temperature (F) 1520 1456 1438 1510 High Pressure Turbine Inlet Temperature (F) 1382 1403 1406 1372 Inlet Pressure (psia) 45.2 52.5 56.3 45.1 Gamma 1.32 1.32 1.32 1.32 Shaft Power (hp) -191.7 -222.4 -238.5 -193.4 Isentropic Efficiency (%) 85.0 85.0 85.0 85.0 Expansion Ratio 2.91 2.76 2.71 2.89 Corrected Flow Rate (lbm/s) 0.85 0.85 0.86 0.86 Recuperator Hot Side dT (F) -113.3 -96.3 -100.8 -97.3 Hot Side Heat Rate (Btu/s) -82.2 -94.1 -100.4 -83.8 Cold Side dT (F) 214.5 207.2 204.7 215.0 Cold Side Heat Rate (Btu/s) 75.8 88.0 94.4 77.5 Heat to Ambient (Btu/s) -6.5 -6.1 -6.0 -6.2 Effectiveness 0.30 0.32 0.33 0.30

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132 Run B4-15 B4-16 Trial 1 2 3 1 Low Pressure Turbine Inlet Temperature (F) 910 959 961 919 Inlet Pressure (psia) 17.7 21.0 22.8 17.9 Gamma 1.34 1.34 1.34 1.34 Shaft Power (hp) -5.2 -16.5 -24.0 -5.0 Isentropic Efficiency (%) 55.1 64.5 66.7 44.3 Expansion Ratio 1.21 1.43 1.55 1.22 Flow Rate (lbm/s) 0.38 0.54 0.62 0.39 Corrected Flow Rate (lbm/s) 0.52 0.62 0.67 0.52 Hot Gas Cooler Gas Side dT (F) -410.0 -390.4 -385.6 -415.0 Gas Side Heat Rate (Btu/s) -79.8 -100.2 -104.7 -83.0 Warm Gas Cooler Gas Side dT (F) -396.1 -407.0 -406.6 -399.2 Water Side dT (F) 10.7 16.2 17.9 10.3 Gas Side Heat Rate (Btu/s) -77.0 -100.9 -106.5 -80.8 Water Side Heat Rate (Btu/s) 59.0 80.1 84.7 61.6 Heat Loss to Ambient (Btu/s) -18.0 -20.8 -21.8 -19.3 Effectiveness 0.70 0.64 0.64 0.70 Cold Gas Cooler Gas Side dT (F) -57.1 -89.0 -93.9 -58.2 Gas Side Heat Rate (Btu/s) -21.4 -37.9 -43.1 -21.7 Overall Performance R 1.11 1.37 1.29 1.12 Power (hp) 67.2 75.5 80.2 66.7 Thermal Efficiency (%) 13.82 13.81 13.89 13.60 Average Water Extraction Rate (gph)2.78 0.96 0.85 3.16 Average Fuel Flow Rate (gph) 9.65 10.85 11.45 9.73

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133 Table C-8. Data from run B4-17. Run B4-17 Trial 1 2 Main Air Inlet Ambient Temperature (F) 86.2 89.2 Ambient Pressure (psia) 14.7 14.7 Ambient Humidity Ratio 0.01770.0180 Mass Flow Rate (lbm/s) 0.00 0.00 Low Pressure Compressor Inlet Temperature (F) 93.1 96.3 Gamma 1.40 1.39 Shaft Power (hp) 9.2 19.9 Isentropic Efficiency (%) 47.6 61.2 Pressure Ratio 1.14 1.33 Flow Rate (lbm/s) 0.62 0.75 Corrected Flow Rate (lbm/s) 0.64 0.78 High Pressure Compressor Inlet Temperature (F) 78.1 84.8 Inlet Pressure (psia) 16.7 19.6 Adiabatic Exit Temperature (F) 330.4 334.9 Gamma 1.38 1.38 Shaft Power (hp) 133.4 150.9 Isentropic Efficiency (%) 72.6 70.9 Pressure Ratio 2.87 2.76 Total Gas Flow Rate (lbm/s) 1.51 1.72 Corrected Flow Rate (lbm/s) 1.35 1.32 Combustor Fuel Flow Rate (lbm/s) 0.020 0.021 Phi 0.46 0.41 Combustion Efficiency 95.7 96.2 Combustor Exit Temperature (F) 1463 1444 High Pressure Turbine Inlet Temperature (F) 1377 1415 Inlet Pressure (psia) 48.0 53.9 Gamma 1.32 1.32 Shaft Power (hp) -204.8 -231.5 Isentropic Efficiency (%) 85.0 85.0 Expansion Ratio 2.87 2.76 Corrected Flow Rate (lbm/s) 0.85 0.86 Recuperator Hot Side dT (F) -73.6 -101.6 Hot Side Heat Rate (Btu/s) -89.8 -98.0 Cold Side dT (F) 218.3 210.0 Cold Side Heat Rate (Btu/s) 83.7 92.1 Heat to Ambient (Btu/s) -6.1 -5.9 Effectiveness 0.32 0.33

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134 Run B4-17 Trial 1 2 Low Pressure Turbine Inlet Temperature (F) 950 963 Inlet Pressure (psia) 19.1 21.9 Gamma 1.34 1.34 Shaft Power (hp) -9.2 -19.9 Isentropic Efficiency (%) 59.2 65.2 Expansion Ratio 1.30 1.49 Flow Rate (lbm/s) 0.45 0.58 Corrected Flow Rate (lbm/s) 0.57 0.65 Hot Gas Cooler Gas Side dT (F) -405.6 -401.0 Gas Side Heat Rate (Btu/s) -96.0 -103.7 Warm Gas Cooler Gas Side dT (F) -433.1 -429.5 Water Side dT (F) 14.1 15.2 Gas Side Heat Rate (Btu/s) -100.8 -108.0 Water Side Heat Rate (Btu/s) 76.2 81.4 Heat Loss to Ambient (Btu/s) -24.6 -26.6 Effectiveness 0.71 0.68 Cold Gas Cooler Gas Side dT (F) -58.6 -79.6 Gas Side Heat Rate (Btu/s) -23.9 -35.3 Overall Performance R 1.42 1.29 Power (hp) 71.4 80.6 Thermal Efficiency (%) 13.84 14.49 Average Water Extraction Rate (gph)2.15 1.40 Average Fuel Flow Rate (gph) 10.24 11.04

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135 LIST OF REFERENCES [1] Gasparovic, N., 1968, The Advantage of Semi-Closed Cycle Gas Turbines for Naval Ship Propulsion, Naval Engineers Journal 80, April 1968, pp. 275-281, 333 [2] Nemec, T. S., 1995, Thermodynamic Design Point Study of a Semi-Closed Recuperated Intercooled Gas Turbine Combined with a Rankine Bottoming Cycle, Masters Thesis, Dept. of Mechanical Engineering, University of Florida. [3] Landon, J. C., 1996, Design and Of f-Design Point Study of Two Regenerative Feedback Gas Turbine Engines for Marine Applications, Masters Thesis, Dept. of Mechanical Engineering, University of Florida. [4] Danias, G. E. 1998, Design and Of f-Design Point Study of Two Regenerative Feedback Gas Turbine Engines for Helicopter Applications, Masters Thesis, Dept. of Mechanical Engineering, University of Florida. [5] MacFarlane, R. S., 1997, A Study of the Impact of Water Extraction on the Regenerative Feedback Turbine Engine Cycle, Masters Thesis, Dept. of Mechanical Engineering, Un iversity of Florida. [6] Muley, N. S., 2002, Effect of Exhaust Gas Recirculation on Thermal NOX Formation Rate in a Gas Turbine Engine, Masters Thesis, Dept. of Mechanical Engineering, University of Florida. [7] Boza, J. J., 2003, Performance of a Semi-Closed Gas Turbine and Absorption Refrigeration Combined Cycle, Masters Thesis, Dept. of Mechanical Engineering, University of Florida. [8] Khan, J., 2006, Design and Optimization of a Distributed Generation System with the Production of Water and Refrigerati on, Doctoral Dissertation, Dept. of Mechanical Engineering, Un iversity of Florida. [9] Lear, W. E., Laganelli, A. L., 1999, H igh Pressure Regenerative Turbine Engine: 21st Century Propulsion, Fina l Test Report for Contra ct No. NAS3-27396. [10] Crittenden, J. F., Lear W. E., Azzazy, M., 1999, Exploratory Design of a Depleted Oxygen Gas Turbine Combustor, Final Re port for Contract No. NAS3-27759. [11] Pringle, D. S., 1972, Testing a Sm all Gas Turbine In the Laboratory Shop, Sawyers Gas Turbine Engineering Handbook Vol. III, Gas Turbine Publications, Inc. [12] Lear, W. E., 2006, Test Report for a Novel Combined Cycle Turbine Engine with Water Harvesting, Test Report for Subcont ract to Phase II SBIR Contract No. A03-037.

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136 [13] Brown, D., 2005, Internal Report fo r US Army Phase II SBIR Contract No. A03037. [14] Gater, R. A., 2002, Engineering ThermodynamicsFoundation Topics Textbook Published by Author. [15] Incropera, F. P., DeWitt, D. P., 2002, Fundamentals of Heat and Mass Transfer 5th Edition, John Wiley & Sons, New York. [16] 1958, Rover Gas TurbineService Literature Rover Gas Turbines Ltd., Solihull Warwickshire England, P ublication No. SP/101/758. [17] Crittenden, J. F., 1999, Dilute, Kine tically Controlled Combustion Efficiency Prediction for Recirculating Semi-C losed Gas Turbine Cycles: A NonDimensional Approach Using the First Damkohler Number as a Parameter, Masters Thesis, Dept. of Mechanical Engi neering, University of Florida, pp. 5657. [18] Holman, J. P. 2001, Experimental Methods for Engineers 7th Edition, McGrawHill, Boston. [19] Volponi, A. J., 1999, Gas Turbine Para meter Corrections, Journal of Engineering for Gas Turbines and Power, 121 Oct. pp. 613-621. [20] Brown, D. Internal Progress Report fo r US Army Phase II SBIR Contract No. A03037.

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137 BIOGRAPHICAL SKETCH Eric Howell received his Bachel or of Science degree cum laude in mechanical engineering from the University of Florida in 2004. His deep-seated fascination with the thermal sciences compelled him to pursue a masters degree at the University of Florida, while working as a Research Assistant in the Ener gy and Gasdynamics Systems Laborat ory. The Author received his Master of Science degree in mechanical engin eering from the University of Florida in 2007.