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https://ntrs.nasa.gov/search.jsp?R=19820008611 2020-03-18T08:55:51+00:00Z
DOE/ NASA/ 003780/ 2 yNASA-CR-165130
.a,
JDA FDR 10086 el
ADVANCED GAS TURBINE (AGT)POWERTRAIN SYSTEMINITIAL DEVELOPMENT PROGRESSREPORT
Research Staff of Ford Motor Company am ,'yf
AiResearch Manufacturing Company of ArizonaDivision of the Garrett Corporation f\ ^F `^`R
August 1980
Prepared forNATIONAL AERONAUTICSAND SPACE ADMINISTRATIONLewis Research CenterCleveland, Ohio 44135Under Contract DEN 3-37Amendment 2
For
U.S. DEPARTMENT O F: ENERGYOffice of Transportation ProgramsDivision of Automotive TechnologyWashington, D.C. 20585
Development
(NASA-CR-165130) ADVANCEL GAS TURBINE (AGT)POYERTRAIN SYSTEM INITIAL DEVELOPMENT REPORTProgress Report, 20 May - 24 Sep. 1979 `FordMotor Co.) 150 p UC AC7/MF A01 CSCL 10B
N82-16485
UnclasG3/44 07993
as u,,, ^Y. 5i. + _ __ ik.R h . :- _v
FOREWORD
This report presents the results of an additional four-month study effort of an Ad-vanced Gas Turbine Powertrain for automotive applications undtsr partial sponsorship ofDOE/NASA Contract Number DEN3-37. The work was performed by the ResearchStaff of the Ford Motor Company and AiResearch Manufacturing Company of Arizo-na, A Division of The Garrett Corporation, as a joint team effort.
The principal investigator for the program was Mr. B. T. Howes of the Fgrd MotorCompany assisted by Mr. E. E. Strain, Mr. R. A. Rackley, and Mr. I R. Kidwell ofAiResearch. Report coordinator was Mr. D. L. Carriere of the Ford Motor Companyand the NASA Program Manager was Mr. Richard P. Geye.
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TABLE OF CONTENTS
rmeSUMMARY..................................................................................................................... iii
1,0 INTRODUCTION., ..................... I .... I ................... ..................
2.0 TECHNICAL DISCUSSION ..................................................................................4
2.1 Master layout : ,efinition ................................................................................4
2,1.1 Powertrain Layout ..................................................................................5
2,1.2 Power Section .........................................................................................8
2.1.2.1 Aerodynamic t-Iowpath ......................................................................8
2.1.2.2 Structural load Path .........................................................................8
2.1.2.3 Rotor D,°, namics .................................................................................. 8
2.1.2.4 Bearings and Seals ............................................................................11
2.1.2.5 Ceramic Turbine Shaft Attachment ....:...........................................16
2.1.2.6 Thermal and Structural Analysis ....................................................21
2.1.2.7 Static Structure; Thermal Analysis ..................................................21
2.1.2.8 Steady State Temperatures ..............................................................23
2.1.2.9 Ceramic Structures ...........................................................................23
2.1.3 Gearbox .................................................................................................34
2.1,3.1 Differential VSTC Gearbox .............................................................34
2.1.3.2 Split Path VSTC C;earnox ...............................................................40
2.1.4 Transmission .............
2.1.5 Controls Analysis ....................... ...........................43
2.I.S.1 Contro! Systems Analysis .................................................................43
2,1.5.2 Controls Model .................................................................................43
2.1.6 Cycle Analysis .......................................................................................43
2,1.7 Vehicle Performance ....................... .........46
2.1,8 Compressor Definition .................................. ........................................49
2, 1.8.1 Compressor System Optimization...................49...............................
2.1.8,2 Initial Impeller Blade Design ..........................................................54
2.1.8.3 Compressor Test Rig ........................................................................59
2.1,9 Regenerator Definition .........................................................................66
TABLE OF CONTENTS (Cont'd)
2 .1.9.1 Regenerator ('-;;,r and Seal Configuration .....................................70,.1.9.2 Regenerv,or Drive System ...............................................................722.1.9.3 Regenerator 'Test Rigs ......................................................................73
2.1,'0 Combustor Definition ........................................................................73
2.1,10.1 Combustor Test Rig ...................................................................... 99
2.1,1 1 'Turbine Definition ......................................................................... 99
2.1,1 1.1 Prclimin; ► ry Investigation .............................................................. 99
2.1.11.2 Stator Design .................................................................................104
22,1.11,2.1 Stator Inlet Flowpath .............................................................. 1042.1.11.2.2 Stator Aerod y namic Design ...................................................... 1122.1.11,2.3 Stator Mechanical Design ...................................................... I I6
2.1,1 1.3 Turbine Rotor Design .................................................................. 116
2.1,11,3.1 C'cr;amic Rotor Aerodynamic Design .........................................1162.1.11,3.2 Ceramic Rotor Mechanical Design2.1,11.3.3 Metallic Rotor Design ...............................................................1262.1.11.3.4 Cold Turbine 'rest Rig ............................................................. 126
2.1,12 Ceramic Rotor Process Definition .................................. ... ... I.,....... 140
2,1.12.1 Matcrials Sclectuar . . ....................................................................... 14U2.1.12.2 Sintered SiC` .................................................................................. 1402.1.12.3 Sintered Si3N4 ................................................................................ 1412.1,12.4 HIPIP Si 1N4 ..................................................................................... 1412.1,12.5 Reaction Sintered SiC" ..................................................................142
LIST OF' TABLES
Table Title pageI Bearing Description-Ball, Split-Inner-Ring, 202 142 Incaloy 903 'tensile Properties (Typical) 213 AGT Performance Comparison 1435 kg (3163 lb.) GVW 464 Reference Conipressor Features 535 Regenerator Pinion and Ring Gear Data 726 Comparison Between Nord VG Combustor and Proposed AG`I' 82
Comhu;tor7 Stator lnlct Flowpath Configura,ion Study 106h ('e-imic Turbinc 'Wheel Properties 1249 Metal l':ori-ine Wheel Properties 135
il
SUMMARY
This report presents the results of an additional four-month in-depth study of an Ad-v,tneed Gas Turbine (AG"l'.) l'owertraain for automotive application. This study was par-tially Supported by the Research Staff of the ford Motor Company in conjunction withA Research Manuftwturin«z Company of Arizona, A Division of The Garrett Corpora-tion, under modifications to DOE/NASA Contract DRN3-37, Results of a prior studyconducted under this contract previously have been reported in "Conceptual DesignStudy of Improved Automotive Gas Turbine Powertraain Final Report,"DUE?/NASA/00374, NASA CR-159580,
't'he principal objective.s of (lie prier study were-,
• fuel economy as minimum of 20 percent l'e'tter than 8.3 km/I 09,6 mpg) for a 154kg (3400 pounds) test: weight baseline vehicle, when tested over the Combined federal[giving Cycle at euli:ssions objectives of 0,41 gni/mile HC, 3,4 gm/mile CO, and 0,4gna/mile NOX
• Acceleration capability of 15 seconds from 0 to 96,5 km/hr (0 to 60 nrph)
• A design capable of reaching as level of development in 5 years (by the year 1983)such that a decision to start a production program could be considered
Additionally, ford requirements for the engine were,
• A fuel economy objective. of 30 to .50 percent be ter than the baseline vehicle
• .A four-second vehicle Start-up distance of 19,5 ni (64 feet)
Prior to initiation of this additional four-month effort, a .request for proposalNASA RFP 3-823389Q for the "Advanced Gas 'turbine Powertrain System Develop-ment Project, (Phase l and Phase 11)" was received, Objectives, as stated in the RFP, areas follows:
• At least a 30 percent iniproveallent in Combined Federal Driving Cycle (CFDC) econ-only (mpg) on the IPA test procedures over that presently predicted for a comparable1984 production vehicle pea ,,vered by as conventional spark-ignition powertrain system(baseline vehicle)
• Gaseous emissions and particulate levels less than the following: NOX014, HC0,41, CO 3A glli/a'aaile and as total particulate level of 0,2 gm/mile using the Salliefuel as used for fuel economy measurements
• Ability to use a variety of alternate fuels
"N
These objectives were adopted for this study conducted during the period of May 20through September 24, 1979,
The major efforts conducted during the study reported heroin were directed towardmore detailed definition and ?analysis of the muster design and performance of Ivey com-ponents of the powertrain concept as, &rived during the prior study, A total "system.design" philosophy was employed whereio distinct recognition was given to the passengervehicle duty cycle over the CFDC, in particular, the amount of time expended at lowpower levels. Therefore, in the "system design" approach, component optimization witsmade primarily at intermediate power levels,
The reference powertraain design (Ril l)) res lAing front these studies is defined its acontinually updated, documented and characterized powertrain design that will serve asthe finalized concept. It consists od' a single shaft regenerated gas, turbine engine utilizin g
ceramic heat section components, coupled to as split differential gearbox with an availablevariaahle stator torque converter (VSTC) and an avadab l,e Ford Integral Overdrive([101)) four-speed ?automatic transtnission, Predicted fuel economy using gasoline fuelover they CyFDC is 0.3 kani/1 (36,0 mpg), which represents a 59 percent irnl^, ovementover the spark-ignition-powered baseline vehicle. Using 0172 fuel, CI^^DC mileage esti-mates are 17,43 kni/1 (41.0 napg). Zero to 96,6 kni/hr .(60 nigh) a a.cceierEatican time is11,9 seconds with as f*our-sex^ond ?acceleration distance of 21,0 m (69 fret) which providesthe AOT ,ichicle with d'riveability characteristics, equal to or better than the baselinevehicle, All perforaaaantc was estimated oat 163,1 m (500 f leet) elevation and 29,40Wll ).
Poring tile: Reference: l Study, shaft speeds of 90,000, 1.00,000 and 118,000 rpm wereevaluated with the 118,000 rpm speed point selected which resulted in the; lowest rotorinertia and hest, rotor response tiaaae. Subsequent design studies on the compressor versusshaft speed indicate that at 118,000 rpm the high inducer Tip Mach number wouldresult in low efficiency and increased technical risk ?associated with inducer leading edgeblade thickness, In addition, studies indicate potential technical risks associsated with the118,000 rpm specie nand tile; predicted third critical speed, For these reasons a 100,000rpnt dcsi,gn sped was selected.
Wfinition or the cerataaic radial turbine rotor has evolved. The radial turbine,fabricated from near-net shape ceramic material using one of several different manufac-turing approaches, is predicted to achieve good part-speed performance while sacrificingrninimuna performance potential at maximum power conditions, The study shows a stat-or vane count, between 1{a and 21 with a rotor blade count between 12 and 14, Stressanalysis indicates as peak rotor stress near the rotor centerline of approximately 206.8MPa (30 ksi) under imaximum power steady state operating conditions.
The coaaaprossor is characterized as a 40- to 50-degree backward swept impeller having12 full blades and 12, spli:tter blaaA vs, A radial diffuser system has been preliminarilyevaluated to achieve: the desired surge; margin and efficiency characteristic over tlaerange of AGT operating conditions.
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Further definition of the variable geometry combustion system and fuel delivery con-cepts wits conducted, Primary and dilution airflow optimization r vats evaluated for thewide range of operating charateteristics to maintain the 1.644'K (2500**) turbine inlettemperature, low emission levels, and desired combustor pressure drop of three -percent,
Steady-state thermal and stress analyses have been conducted on the structural com-ponents of the RPD, Results indicate that with internal insulation, component stressesare acceptable over the operating range, Design iterations will continue with selectedceramic manufa►cturors to further address such problems as ceramic-to-ceramic and ce-rica, ic-to-metallic interfaces, n anufacturability, part integrity and reliability, and thesa ► btleties of near-net shatix c vra ►w4, c pr^,Aucibilit,y,
Definition of the regenerator care, seals and drive system was accomplished, An ex-truded matrix cord with at hydraulic diameter of 0,508 mm (0,020 inch) has been select..-cad for the engine, Three inner :weal configurations were examined and evaluated withrespect to overa ll system of i'i6ency, An equa l now area split design (balanced h ea t ex.,changer) design provides the Best regenerator effectiveness and pressure drop combin g►-tion throughout. the engine operating range, A regenerator drive iystetn, similar to theFo,rd 707 system, wits analyzed and recommended 1`or detailed design,
Two split path differential gearbox design configurations were evaluated, In one, theV STC is required to operate briefly under it negative speed ratio up to it vehicle: speed of6A to 9,7 kni/hr (4 to 0 naph), The second split path design configuration does notnecessitate operation of the VSTC in the negative speed ratio, Determination of therecommended approach will be made following more detailed ►napping, of the variablestator torque converter characteristics,
The control system evaluated for the pow rtrain is a full authority electronic system,A inicropro :essor in the control package accepts the sensor signals, reacts, and drives thefuel control and variable geometry acCjators to operate the powertrain in a responsiveand fuel efficient manner. Automatic start sequencing and protective features have beendefined to provide to safe basic philosophy and the same "feel" and drive input, is atconventionally powered vehicle,.
Th -. ceramic radial rotor i's the longest.-lead, highest-risk ceramic component, Designstudies indicate sintered SON4 and SiC, reaction sintered SiC (RSSiQ and hot isostaticpressed (HIP) Si3N4 have the potential for meeting the project requirements within thedevelopment timetable,
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1.0 lKYRODUCTION
This report, submitted by the Turbine Research Department of the Hord Motor Com-pany in conjunction with the AiResearch Manufacturing Company of Arizona, a Divi-sion of The Garrett Corporation, presents the results of a recently completed four-monthin-depth study as specified by modifications to DOE/NASA Contract DEN3-37, to con-tinue definition and analysis of an advanced gas turbine (AGT) powertrair, system forautomotive application. The work reported herein covers the period of May 20, 1979through September 24, 1979, and is considered an extension of the design study -onduct-ed under DOE/NASA Contract ;DEN3-37 previously reported as "Conceptual DesignStudy of Improved Automotive Gas Turbine Powertrain — Final Report,"DOE/NASA/0037-79/1, NASA CR-159580, Reference (l)
The principal objectives of the reference study were:
a Fuel economy a min.imuna of''20 percent better than 83 km/l (19.6 mpg) for a 1542kg (3400 pounds) test weight baseline vehicle, when tested over the CFDC at emis-sions objectives of 0,41 gm/mile HC, 3A gm/mile CO, and 0,4 gm/mile NOx
• Acceleration capability of 15 seconds from 0 to 96{5 km/hr (0 to 60 mph)
• A design capable of reaching a level of development in 5 years (by the year 1983)such that a decision to start a production program could be considered
Additionally, Ford requirements for the engine were;
• A fuel economy objective of 30 to 50 percent better than the baseline vehicle
• A 4-second vehicle start-up distance of 19.5 m (64 feet)
Investigation during the initial study included some twenty-two different engine con-figurations and nineteen candidate transmission concepts (three for a two-shaft configur-ation), Selection cr'teria were established and used to evaluate the concepts. These crite-ria were weighed and appear as follows in order of rank; fuel economy, manufacturingcost, technical risk, installation considerations, and development and growth potential,As reported in Reference 1, AiResearch Document 31-3529, a single shaft regenerativeengine with a centrifugal compressor with variable inlet guide vanes (VIGV), variablegeometry combustor and single stage ceramic radial turbine rotor coupled to a differen-tial gearbox containing an available VSTC and an available Ford FIOD four-speed auto-matic transmission, was recommended for further study.
Prior to initiation of the continued study, a Request For Proposal----- NASA RFP 3-823389Q --- for the "Advanced Gas Turbine Powertrain System Development Project(Phase I and Phase 11)" was received. Objectives, as stated in the RFP, are as follows:
• At least a 30 percent improvement in CFDC economy (mpg) on EPA test proceduresover that presently predicted for a cornpatible 1484-preduction vehicle powered by aconventional spark - ignition powertrain system (baseline vehicle)
• Gaseous emissionti and particulate levels less than the following; NO X = 0,4, HC =0,41, CO 3A gm/mile and a total particulate level of 0,2 gin/mile using the samefuel its used fir NO economy measurements
t, 1.
• Ability to use a variety of alternate fuels
Therefore, all performance comparisons and supporting analysis is in response to thegoals stated in the .RUT, In addition, the RFP identified the following nomenclature andinterpretative definition for the powertrain program, which has been adopted as follows-,
Reference Powe rtrain Desig.ri (RI'D) . Continually updated, documented and char-acte tri ed powertrain design that will serve as the finalized concept
• Mod 1 Engine A ;stepping stone engine incorporating the advancing technologiesbeing derived from the con.;wnent development activities
The definition and analysis efforts of this study were based on the recommended con-ce.ptual design layout as presented in Reference 1, The principal tasks included;
• Continued definition of the hot section of the engine and ceramic components includ-ing aerodynamic and stress analysis of the turbine exhaust diffuser and regenera-tor/turbine ducting; and analysis of potential regenerator configurations, includingcore and seal geometries and regenerator drive and housings
• Continue to analyze performance considering effects of the engine definition itera-tions, including determining the effects of differential and sprit-path variable statortorque converter /gearbox/transmission concepts on combined federal driving cycleperformance; and effects of control system relative to powertrain/vehicle transientcharacteristic;c
• Continue aerodynamic and stress definition of the ceramic turbine wheel concepts anddefinition of metallic turbine wheel concept including preliminary specifications for aturbine test rig and a rotor dynamic simulator to permit suitable testing of the result-ing designs
• Continued definition of the compressor concepts for aerodynamic stress, and materialconsiderations including VIGVs and effects of impeller/VIGV interaction; and prelim-inary specifications for the compressor test rig and VIGV rig
• Continue definition of alternate combustor configurations and analysis of relative per-formance including preliminary definition of a baseline variable geometry combustorand preliminary specification of a combustor test rig
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PHASE L, L-OROVF<U CAS TVRBINE JUNE JULY AUGUST SEPTEMBER OCTOBER
ASK II1 ,C, 1 'MASTER LAYOUT DEFINITION
DES'y GN UPDATE
CYCLE ANALYSIS
CERAMIC STRUCTU,AE
FOIL BEARING
CONTROLS AND ACCESSORIES
REGENERATOR DEFINITION
TA$9 ltl,C.2 TURDINE DSr--:TION
AERO DESIGN CERAMIC
STRESS EVAL CERAMIC
3ATL EVAL CERAMIC
.HERO DESIGN 3ETALLTC
STRESS EVAL METALLIC
`"ATL SELECTION ,METALLIC
EXHAUST DUCTING
PRELIM TEST RXG DESIGN (COLD TURBINE)
PRELIM PEST RIG DESIGN (ROTOR DYNAMICS)
TASK III.C.3 C044PRESSOR OSFINITION
AEROCYNAMIC DESIGN
S'TRES5 EVALUATION
SELECTION OF vt,rL CONSIDERATION
PRELIM TEST RIG DES1vGN CONCEPT
TASK III;C.4 COMBUSTOR DEFINITION
EVAL OF ALTERNATE CONCEPTS
COWT DESIGN OF VARIABLE GEOM
PRELIM TEST RIG DESIGN CONCEPTS
TASK III.D CERAMIC TUABINE WHEEL PROC. DEF,
MFG PROCESS INVESTIGATION
EVAL OF NET SHAPE PROCESS'
F.VAS+ OF HIPPING
Figure 1. Program Master Schedule.
• Continued investigation of silicon nitride and silicon warbide ceramic radial turbinewheel net shape fabrication processes to attain higher strength including hot isostaticpressing alternatives
Figure I shows the program master schedule,
A total "system design' s philosophy was employed during the definition and analysisefforts, Through an iterative design process a reference powertrain design is being estab-lished, The philosophy of "system design" incorporates a distinct recognition of a passen-ger vehicle duty cycle over the CDFC and, in particular, ([re amount of time expended atlow power levels„ Therefore, in the "system design" approach, component optimization ismade at intermediate (off design) power levels and not at maximum power (designpoint) as'is typical of normal gas turbine design techniques, Basic guidelines used inarriving at he reference powertrain design included;
• Performance . Enviromental Impact. -- Consumer Merketability
Technology demonstration for a go/no-go production enginering decision as early as1983
• Demonstration of all vehicle/perfortraswe aspects by 1984.
» Probability of success and assessment of technical risk
This powertrain design is identical in concept to the powertrain concept of Reference 1differing only in the details of key components and subsequent refinements to the per-formance predictions, It consists of a single shaft regeneratod gas turbine engine utiliz-ing ceramic hot section components, coupled to a split differential .gearbox with an avail-able VSTC and an available Nord kIOD four-speed automatic transinission. Controlsfeature accurate microprocessor full authority digital technology, integrating all power-train control functions into one master control system, stressing driveability and integrat-ed performance as related to consumer acceptance, The R.PD package has the flexibilityof providing either front- or rear-wheel drive,
The following sections discuss the continued definition of key components and subsys-tems,
2.0 TECHNICAL DISCUSSION
The following paragraphs describe the continuing definition of the Advanced Gas Tur-bine (AGT) and key comronents,
2.1 Faster Layout Definition
Master layout. definition and refinement has included;
4
• Powertrain• Power Section• Gearbox• Transmission• Controls Analysis• Cycle Analysis• Vehicle Performance• Compressor Definition• Regenerator Definition• Combustor Definition• Turbine Definition• Ceramic Rotor Process Definition
2, 1.1 Powertrsin Layout
As shown in Figure 2 the reference powertrain design (RPD) addressed in this studyincorporates a power section containing the single-shaft ceramic gas turbine and relatedceramic hot, section components, a variable, speed differential variable stator torque con-verter (VSTC) gearbox, and a Ford integral overdrive (FIOD) automatic 4-speed trans-mission,
Due to inherent characteristics of the differential gearbox, design selected, the VSTCis required to operate briefly under a negative speed during the first 6,4 to 9,7 km/h (4to 6 mph) of vehicle speed, Specific VSTC characteristics under negative Speed ratioscould not be confirmed during the study. Consequently, a second powertrain configura-tion with a variable speed ,gearbox not requiring negative speed ratios in the VSTC isbeing investigated.
Prelimimary analysis has indicated that either configuration meets the performancegoals.
Both powertrain designs are approximately 1443.8 mm (57,0 inches) long and weigh237,2 kg (523 pounds) exclusive of conventional automotive accessories, such as powersteering pump, alternator, and starter, The power section is approximately 552,45 mm(21,75 inches) in diameter by 5203 mm k20,5 inches) long, The variable speed gearboxand 'VSTC sections are approximately 698,5 mm (27,5 inchos) wide by 604.4 mm (24,0inches) high by 279,4 mm (11,0 inches) long including accessory mounts, The MODtransmission is approximately 330,2 mm (13 inches) wide by 330,2 (13 inches) high by647,7 .mm (25,5`inches) long, It is anticipated that significant weight reduction can beobtained during detail design,
Figure 3 shows a powertrain utilizing a split path gearbox that does not require opera..lion of the VSTC under negative speed ratio,
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ORIGINAL PAGE 13
OF POOR QUALM
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ORIGINAL PAGE ISOF POOR QUALITY
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2,1.2 Power Section
The AGT ix)wcr scctiorl inctirix►rittes t ltirger 4 ,erunric rcg iicrutt ►r, ceritntie hot soctioncontilxrilents, improved ttcrodyrltinlic Ilowhuths, improved structtnriil load Millis, find ithalted t"rorit rtat!trncritlot , cover in coil Iparison to (lie Reference I coal"igurtttio ti=
2.1.241 Aerodynamic Fiowpath
The A6T crittiiic itero(ytiittllic` 1'fowllirtlt is shown iii 1^igtire 4, Air enters the conipres-
sor inlet i^icnuni tit tittit ► ii , I'lows tllrotigh the iiiiet guide viuies into thert►dinl compressor C.D , through the compressor di1"1"tiser Q4) , itrid into tin mintilitt"space Q-) surromiticd by the A611" miter housing ^(► ("onii►ressor discharge :tit`flows lion) the untililar s xlee nroti rid the rtegenermor drive, gear ittlti toilet'
into the Ili tti liresstu °c' sidt^ tit" t1i4 re eiierlttor ii , otit tit` ilia replierittor mid dtivictilhrtiuKli the I'lciw seiutriit±►i` licius rll► to tile cclitt111tstcrr It will the tlririiary ports
,md dihitiuti ports 1 , Duct is stipplicd to the cc►ttthtistor tltrottgh the 1"uel aiol-101 rte , and con'thtistioti lrrkdticts lire (tirevied iirotittcl ilia coiiillustor t ► tu'l'lethrough the tnrhilw i ot'llc 1 ^ , itrid into lite radial turb ine Qi t , 'r4rrliine dis-
chlirge glues tire directed r,tWit►lly to it illenuiit ^I ) lkild thr otigh the low -presstire shiesof the regeocrlit it \ ` to '..e ttirbille xiiittist
j
2,1.2,2 Structural Load Path
As shown iti Vilture 5, AOT pressure,i:,o it,iitinicrit is lirovitled by the AO T cutler he►us-
nlicl repwileriittli' cover" , 'l"he rergelieM'iYtlir' Sellt 11rt`ltltrtiS ^ , itriti pert►61 11ttrMe iorces are relicted throogh the flow sellitrntor hokising 5^ to litisses
ef
C ► ttnmid lirt►tik;l lt► the ct►iii lr^c,sclr inlet liotrsiri itricl n ittc nit►uiits oil the mating,I, ^ge,'Ir' , ox. AercxtyrumliQ loi tcl, oil the Ilow cellar":tlor howshi t (' = lire relinedtlirottgh -die sttiiie path, C ol -lillu "slor lircltul(l S) is trltiisolitted throtigh the colllbw fortiller (`t , to the tilrhiiic shroud /ncitl.lc It) , laid roil► (ticsttitic hos'ss, ^? its thepresstirc lentils= Acrotipnttwic forces 11 irrc reacted iltrokigh this slime It ►:tt plith.
2,1.2;3 Rotor Dynamics
OtIritig file Referellee 1 sitidy, skirl"t speeds of 90,000, 100,000 tind 1 i 8,0(K) rlliii `ertcvit wMed with Ow l 18,4'00 ri ►ill speed point selected which res llicd ill the lowest rotorincrthk ttrid hest rotor, rc.,Sponscs tu l le " subsecliielit dcsigii studies oil tiie comp ressor versusshaft speed hive indieated that, alt 118,000 rpill, the high inducer tiii Mitch Nitriiher
ccoold result ill low el"l'iQics tiey wid iiicsrcascd techiiical risk itssioc"inted with ivdticer lead.inn edge hlitde thickiicss " lti lidditioll, studies indicated polcloiill teehiktil risk, iis socs ial-ed with the 118,000 rplii spy ed titid the predictett third critical s peed " Vor These reitsoils ti
100,000 rpm desigii speed w,,is selected, Vill thcl, lillirlysis of the AOT cogine cotifigtirit-tioll ttpiit indicated iilstti"1"icient nliiri in ttetwecrt etigilw rotor ojvrmiti)z speed and Sire
3SOR?.LrINUM
Figure 4. Power Section Aerodynamic Plowpath.
FIOU(• S. PoW*lf $*Ctlon Structural Load Paths,
dieted third 41,66eaal speed, To correct this Imential problem, the rotor dynamic analysiswas continued a► tttl changes were incorporated, The rotor shaft was shortened 9,5 min(0375 inch), tile aluminum impeller forward face wa s piloted on the drive shaft insteadof as curvic', coupling, tatld 0%, foil bearing diameter was increased 2,54 mm (0,1 Inch),Combining all modifications and assuming that the gear mid bearing had 4,3117 N /HintOISAX) pound/inch) stiffness and the foil bearing 700 N /mm (4,000 p ound f inch) stiff-ness, the criticaal speeds were calculated its followst.
Critical Sjwedr: r m W-Ncri tiun
1 4,090 Metallic turbineAM) Metallic turbine
;t 1 34, 700 Walk turbine4 208,000 Metallic turbine1 .I'M Ceramic turbine,h 8,260 Ceramic turbine
111 2100 Ceramic turbine4 ?07,100 Ceramic turbine
The third critical sp ell had it marlin of .t4,7 percent over 100 percent speed (100,000rpnl) for the metallic turbine and ,52) percent for the ceramic: turbine, which is consis-tent with Ford wid AiRtcsea rch experience,,
The following bearing loaads were caalculaaWd for as normal CO eccentricity of 0,0127nital MOMS inch).
Absolute Hearing roads, N(Pounds)
rpm C'ear End Hearing Fo l iJearing Mks+eription
1001000 182,4 (41) 13.3 GO Metallic turbine100 1 000 100,1 (36) 8,9 (1) Ceramic turbine
The effect, of the Illalnetaary geaarset oil rotor dytaatntics could taot be satisfactorily re-,solved analytically dale to tile, wide divergence of data and experience, Geacrset influencefactors can the AOT rotor nivist, he estaablistied by actual test in at rotor dynamics test rig,
2.1.2.4 Bearings and Seals
The high speed ball 1wtiring and shaft seal aarmngeawnt currently under evaluation isdepicted in Vigurc 6,
Nearing aanaalys s wits based tail load and speed profiles expected over tine 160,934 kill(1(10,010 'mile) vehicle operaationaal lift, Oil ilia basis cal C"FDC, the high speed bearingwould he operated for 31+90 hours at speeds between 55,000 and 84,000 rpm, The bear-
1
Figure G. Nigh Speed Ball Bearing and Shaft Seel.
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TABLE 1
BEARING DESCRIPTION BALL, SPLIT-INNER -RING, 202 GRADE ABEC 5Inner Ring Outer Ring
Material SAE 52100 STL --- RC 58- Material SAE 52100 STL — RC 5864 64Bore 15 mm (0,5904-0,5906 O D 1,3778-1,3780 (35 mm)Width 13,335. 13,589 mm (0,525.0,535) Width 0,4281-0,4331 (11 mm)Race Depth 24 Min, % Ball/Roller Dia, Race Depth 16 Min, % Ball/Roller Dia,Race Curvature — 51,6-52,4% Race Curvature — 50,6 -51,4%Bali/Roller Dia, Ball/Roller Dia.Separator Pilot Land Separator Pilot LandTo Groove Runout TIR To Groove Runout 0,0127 (0.0005) TIR
Separator Rolling ElementsMaterial — SAE 4340 STL _-- RC 34-38 Material -- SAE 52100 STL ----, RC 584
64Silver Plated 0,0005-0,0015Thick Per AMS 2410
Construction MachinedAssembly One-PiecePiloting Surface — Outer Ring LandsPilot Clearance 0,264.0,406 (0,010-0,016)Total AXIAL y A Clearance of0,198 (0,0078) Max Under 1 Lb GaugeLoad
Elements Per Row 10Element Dia, 6,35 mm (0,25 Inch)Element Length
ClosuresNumber
MaterialConstruction
Commercial Pack
Military Spares PackPreservative — MIL-L-6085 See Note 3
See Notes 1 and 3
AiResearch Part Number 3822082.1 -4
14
a
THRUST METALWASHER RING
AP X 3.45 N/cm2 (5 PSI)
TEMP-394K (2500F)
U.UU /bmm GAP
(0.0003 IN )
Figure 8. Floating Ring Seal.
ii
i
ing would operate at 55,000 to 56,000 rpm for 58 percent of` the pra ,,,ected 160,934 kr(100,000 stile) vehicle life. Bearing lubrication and cooling would be provided by autotnotive transmission fluid (A.TF) jetted into the shaft center and delivered cenlrifugallto the ball bearing split 'inner race,
Using a 15 mm (O.S9 inch) 202 series bearing, the ball bearing concept provides 212hours BI life operating at 1,00,000 rpat and an aerodynamic thrust load of 1134.3 l(255 pounds), This thrust load corresponds to design conditions of 302,7K (85'F) a ►rabicant temperature anti I 521 in (500 feet) altitude, The bearing would provide; 135 hour131 life at the maximum thrust load of 1.334,5 N (300 pounds), corresponding to loiambient tenlpe.rature conditions, In either case:, the bearing appears adequate i'or n1,"xmunt load and speed conditions, since nnly 12 hours of predicted vehicle engine operation over 160,934 km (100,000 miles) will be at 100,000 rpm and maximum thrust,
The hearing design is shown in Figure 7, The bearing specifications are given in Table
The; high speed rotor shaft seal between the gearbox and compressor (Figure= 8) is adouble floating ring design, Each seal consists of a carbon ring that provides f ace andbore scaling surfacesx A metallic ring is shrunk on the carbon OD to provide xrlechanicaxlstrength„ A spring is placed between the two carbon rings to preload face walling sur-faces, Additional sealing force is provided by a purge air st,pply between the seals. Th{air purge feature also prevents any po;isible oil leakage into the compressor inlet,
As a result of changes required to provide adequate margin between operating andthird critical speeds, the fail journal bearing between the turbine and compressor hasbeet, increased in diameter 2,54 mm (0,1 inch) to 34,29 mail (1,350 inches), It 11,18 beendetermined that bearing axial length of 273 mn't (1,075 inch) is adequate for AOTapplication, Seven overlapping toils of a taominal 0,1.27 nani (0,005 inch) thickness arerecommended for the initial design pending final rotor analysis and design,
2.1.2.5 Ceramic Turbine Shaft Attachment
Two methods of attaching the AGT ceramic turbine wheel to a metallic shaft havebeen selected for consideration, The shrink-fit concept shown in Figure 9 involves a me-tallic sleeve shrunk over a ceramic turbine stub shaft, The second concept, shown inFigure 10, incorporates a curvic coupling joint between the ceramic turbine and metalliccompressor, The curvic coupling concept has been demonstrated on the Ford 820 turbineattatchment, but involves difficult ceramic machining and high local stresses,
The shrink-.fit concept appears to offer, the most reliable, and ultimately, the lowestcost attachment for the ACTT cera,,raic turbine, Analysis of the concept consideredsintered silicon nitride and sintered silicon carbide turbine wheels with thermal expan-sion characteristics as shown inFigure 11, The IN903 metallic sleeve material was se-
a to
FOIL SEARING
Figure 9. Ceramic Turbine ltlheN Sleeve Attachment.
17
0
18
5
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SINTERED SILICON CARBIDE
E 406.8 x ^ 0 6 kPa(59 x 10 PSI)
E 299.9 x 166 kPa(43.5 x 10 PS I.)
SINTERED
SILICON NITRIDE
E - ELASTIC MODULUS255-922K (0 - 1,2000F)
00 200 400 600 800 1000 1200 1400 1600
0 200 400 600 800 1000 1200
TEMPERATURE
Figure 11. Mean Thermal Expansion CoefficientTurbine Candidate Ceramics.
iq
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-^.^ w »w \. \ w \^ ^^^ . .
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TEMPERATURE
X
Figure 12. Mean Therma l Expansion Coef ficien t, IN903
2 ` ]
lected for low thermal expansion charactcristics and the physical properties as shown inTable 2 and Figure 12, A nominal 0,051 mm (0002 inch) room temperature interfer-ence between the ceramic stub shaft and mtallic sleeve was selected to permit assemblyat 888,5K (I i 40° F ), Analytical evaluation of the shrink-fit concept is in process to de-termine surface stress distribution,
TABLE 2. INCALOY 903 TENSILE PROPERTIES (TYPICAL).
`GYP «ULT R.A.Temp, K (°F) 105 kPa (105 psi) 105 We (105 psi) (96)
294(70) —11,03 (1.6)
13.10 (1',9) 40922 (1200) 8,96 (1, 3)
WOO (1,45) 55
2.1.2.6 Thermal and Structural Analysis
Thermal grids wer p, developed for an AGT power section structure and rotating group.Results were used to generate boundary conditions, and thermal and stress grids for thestructure, These results were then used in the thermal and stress analyses.
Structure steady-state thermal and stress analyses were conducted at maximum powerand idle power conditions, with and without thermal insulation,.
Similarly, rotating group temperatures were determined for 1644K (2500°F) and1422K (2100'i") turbine inlet temperatures, i.e„ ceramic turbine and dual alloy metalturbine inlet temperatures, respectively,
2.1.2.7 Static Structure Thermal Analysis
Figure 13 illustrates a typical thermal grid network used in static structure analyses.Housings modeled included the intake plenum, compressor diffuser, outer die cast hous-;Y, turbine housing, turbine shroud and diffuser, turbine discharge housing, turbine
back shroud, combustor baffle and transition housings, and foil bearing housing. Figure13 includes fibrous high temperature insulation in the cavities between the combustorbaffle and turbine back shroud, and between the turbine discharge housing and compres-sor diffuser duets.
The models incorporated appropriate boundary conditions, as required, for the insulat-ed and uninsulated analyses, -including gas temperatures and temperature gradients, flowrate distributions, heat transfer coefficients, and free convection and radiation betweenhousings without insulation for both maximum power and idle conditions. In addition,heat generated by the foil bearing for the two power conditions wa,,; shared equally in themodel by the bearing housing and journal.
21
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ORIGINAL PAGE 18OF POOR QUALITY
22
2.1.2.8 Steady State Temperatures
The anmlyses revealed that the uninsulated structure would not be satisfactory, Theuse of insulation is necessary to reduce thermal loads into the turbine rotor and foilbearing housing, and reduce thermal gradients across the turbine exhaust housing andLAS flow separator housing, fior these reasons only the insulated configuration, Figures14 and 15, is presented, These figures tare thermal maps for the maximum power andidle power conditions, respectively,
Maxinauna temperatures throughout tii+-, structure foe the tw ,.) steady state operatingconditions were found to be tat acceptable levels for the materials used, except for thecompressor inlet and diffuts housing, Based on these results the use of cast iron forcompressor inlet and diffuser housings is currently being investigated,
Rotating group thermal aaaaps for the 1644K (2500 1 1) turbine inlet temperature aregiven in figures 16 through 19,
in addition, prelinainar y tesults have been obtained for interaction between thestationaary and rotating parts of the foil bearing, These results are shown in Figure 20,
2.1.2.9 !ceramic Structures
Preliminary 2-dimensional stress and deflection analyses wem completed for the se-lected Rill) major cerauaaie structural components (see Figure 21), No major structuralor thermal problems were revealed except for the turbine shroud (Figure 22) which hadsignificant naechaatic aa) atnRl thermal deflection. 'T`his deflection would caause vaari , tion inthe turbine-to-shroud clearance and subsequent performance variation, Therefore, thehousing was modified to the configuration shown in Figure 23,
Maximum power rotating group clearances have been calculated for the modified tur-bine shroud duct. with an aalunainum and cast iron comiaressor inlet housing and acre givenin Figures 24 and 2S, respectively,
l?volution of ` ceramic components to the configuration shown in Figure 26, as coni-pared to the reference configuration, has been the result. of thermal and structural analy-ses, gas path aerodyna► mice improvements, ceramic material technology litnitations, re-generator growth, and engine assembly considerations,
Further evolution of existing geometries is anticipated, In particular, this will involvethe flow separator hoaasing, combustor outer discharge housing, combustor baffle turbinenozzle, turbine back shroud, turbine outer and inner diffuser housings, and combustor,The same factors ,is, discussed previously will govern these changes,
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(1941) TIT 1644K (2500°F)i
13201290
(1917) f
(1862)
1,264(1015)-----
1026 686 (775 )
I^ 1320 1246 (1388)(901, )(1, 899) 1,2 97 (1.783)
(1876)
:w
Figure 16. Thermal Map of Turbine Rotor.
46 5472 462 (378) 412(391) (373) (282)
41.2635 (283 )(683)588
(599) 445 (341)
686 612 615(775 ) (643) (648
1026 -`(1388)
1045(1422)
800(981)305
(990)
637(688)652M5)
TIT 1644K (2500°F)
Figure 17. Thermal Map of Foil Bearing.
37
1
445(342)
44 a(342 )
431.(315 )
TIT 1644K (25000F)
414(2 86 )
381
414 390 (226)
(285) (243 )407 386385(234)
411(274) (235)
(281.)403 333
(256) /(140)
333 -(140) 354 —.
(178) 389(241)
363 411
_ (201) ---- - - --- (194) — (280)
TIT=1644K (25000F)
Figure 19. Preliminary Thermal Map of Shaft (Output End).
29
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1212(1723 )
442,-1(337)
411(281)
419(2 94 )
1043(1418)
(341)
Figure 20. Foil Bearing — Junction Steady-State Temperatures At Maximum Power,
30
CERAMIC COMPONENTS
Figure 21. Preliminary AGT Engine.
31
a
i 1
0.1854 mm (0.0073 IN ) MHCH & THERMAL0.356 mm (0.0140 ZN ) MECH ONLY
Figure 22. Initial Design Turbine Shroud Deflection.
32
Figure 23. Modified Turbihe Shroud.
+0.2286=,U (+0-009 1uvt,'U ONy 44 -0.44131iM (-0.0095 IN )
MECH & THERMAL
33
2,1.3 Gearbox
During evaluation of RPID over the CIiDC with the differential gearbox, the VSTC isrequired to operate briefly under a negative speed ratio during the first 6x4 to 9,7 km jh(4 to 6 mph) of vehicle speed, pecific VSTC characteristics under negative speed ratioscould not be confirmed during the study, As a consequence, a second gearbox configura-tion with a variable speed gearbox not requiring negative speed ratios is being investigat-ed, Preliminary analysis has indicated that either configuration meets the performancegoals,
2.1.3.1 Differential VSTC Gearbox
The differential VSTC gearbox is shown in Figure 27, A schema ,'s`? of the final recom-mended configuration is shown in Figure 28 and defines the gear sizes, tooth numbers,ratios, and range of speeds for the various components,
The following featrUres reflect the current design;
e Carrier planet center distance was decreased to as small a dimension as possible todecrease planet bearing centrifugal loads and provide an increased ratio down to theFIOD input, The first mesh ratio was changed to 11,2727 from 1.1.00 to eliminatehigh common factors in this critical mesh, Planet clearance was maintained
e Diametral pitch of` the first mesh was increased to 28,7 to allow sufficient wall thick-ness under the high speed pinion teeth and enough inside diameter to fit on the mainr/)tor shaft with the 202 series ball bearing, The 283 diametral pitch was also selectedto achieve the proper planet center distance with the compound planet and ring gear
• A tooth tip compressive stress problem with the 14-tooth planet at, low vehicle speedsand high torque ratios was solved by a change to 16 teeth, This changed the `R' ratio(fixed carrier compound ratio) from 47,8571 to 47,5114
These small ratio changes were largely compensated for by adjustment of the revers-ing `X' ratio on the torque converter input side, This is a fixed carrier planetary withthe reduction ratio changed to 2,2143 front 2,2449, The sun gear was maintainedlarge enough to clear the modulating clutch outside diameter
Pace width is critical only at the small 16-tooth planet gear where 14.5 mm (0,57inch) maximum was used. This is approximately one pitch diametee, The high speedpinion face width is 6;35 mrn (0,25 inch)
e High loading and the need for extreme light weight dictates carburized AISI 9310material for the rotating carrier planet gears, In addition, the high speed pinion andsmall ring gear will be carburized steel, All other gears may be induction hardened orthrough hardened AISI; 4340 or equivalent
34
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36
Figure 26. AGT Power Section.
37
Figure 27, Differential VSTC Gearbox.
38
2540 TO 8870 RPM
248T
x L s
1147 TO 4007 RPM
248T
X - 2.2143 (JOIN
67Ty112T
16T VSTC
-2103 TO +1829 wcalmisla RPM
OUT113T
55,000 22Tw l 0 ,000
D p 28.7
D p = 28.01 1 - 1 l;,
wo 0-3900 RPMZ = 1.3672850 RPM MAXFIODINPUT SHAFT
R = COMPOUND RATIO 113/22 X 148/16 = 47.5114Q RING GEAR/PINION 248/22 - 11.2727X ;= REVERSING RATIO 248/112 = 2.2143
NEGATIVE SPEED RATIO (w 0 = 0) = —0.5253SPEED RATIO (VSTC) SR - w A
e n Bout1 = R (Q+1) SR + QX (R+1)
woR (SR) + QX
Dp a DIMETRAL PITCH
Figure 28, Differential VSTC Gearbox.
19
• Thu `Z' ratio (gearset between differential output and MOD input) was designed its a15 diametral pitch helical. mesh with a helix angle of approximately 23 degrees, Agear face width of 13, 2. inin (0,52 inch assures an overlap ratio of exactly one, pre-cluding axial load shift through the plane of action
• Critical bearings were analyzed for 100-hour test conditions at the most severe load-ings, Needle bearings supporting the planets will provide sufficient life at the highestcarrier speed of 3937 rpni, Hollowed shafts and holes in the planet gears will be uti-
lized for weight reduction
• The `Z' ratio gears, one oft the carrier and one on the FIOD input shaft, are adequate-ly supported on ball bearings that react to the helical gear thrust
2.1.3.2 Split Path VSTG Gearbox
The alternate split path VSTC gearbox is shown in Figure 29, A schematic of theconfiguration is shown in Figure 30, The split path design differs from the differential.design primarily in the path the VSTC power is returned to the output, shaft of thegearbox, In the split path design, a portion of the input power is delivered to the singlecompound planetary ring gear, through the `X' ratio reversing gear set, into the VSTCand to the output shalt via a clutch, The balance of the input power is transmitteddirectly into the planetary carrier and output shaft.
In the differential VSTC design, input power is also split, into two paths, One powerpath is through one of two ring gears meshing with the compound planetary, through theVSTC, a clutch, and then back into the compound planetary via the second of the tworing gears. The balance of the input power is transmitted into the planetary carrier andoutput. shaft as in the differential design,
The ;split path design incorporates a compound planetary gearset with a compoundratio of 14,52;1 as compared to a ratio of 47,5;1 for the differential gearbox. The lowernumerical compound ratio results from use of larger compound planet gears with corre-sponding lower stresses, The first mesh of high-speed pinion and planet gears was madeidentical to the initial differential gearbox design to facilitate early performance compar-isons of the two gearbox designs. Optimization of the split path design would ultimatelyresult in equalizing; the compound planetary ratios so that the first mesh would be ap-proximately 4;1 instead of the 5:1 shown on the schematic, This change would reduce
the carrier diameter and gear Stresses,
2.1.4 Transmission
A modified I IOD automatic 4 -speed transmission with the variable speed gearboxprovides a close approximation of an infinitely variable transmission required for single-shaft gas turbine veh cid engines. Modifications to the ROD would include;
40
si
Figure 29. Split Path VSTC Gearbox.
41
a
i 1450 TO 5376 RPM219T
3786 TO 6885 X = 1.2807
^ RPM
24T +A'cin183T
i T 171T
J63T 0 to 2930
100T_ 0°out Q
55,000 - 100,000T ,-0T0TO 2930 VSTC
w. w
Z = 1.1018
R = COMPOUND RATIO = 100/20 X 183/63 14.5238SPEED RATIO (VSTC) SR w /w
oout °inX REVERSING RATIO = 219/171 1.2807
w^R (1 + X/S R ) + 1
Figure 30. Split Path VSTC Gearbox.
42
g
• Nixed stator torque converter removal
• Transmission bell Dousing removal and adaptation to gearbox case
• Transmission control and shift logic integration with AGT control microprocessor
• Gearbox and FIOD lubrication systems integration
The F IOD transmission is a fully developed production transmission currentl y used in1980 Ford vehicles, No other modifications would be required,
2.1.5 Controls Analysis
2.1.5.1 Control System Analysis
The control system defined for the AGT is shown schematically by Figure 31, Thesystem was programmed for dynamic computer studies of the powertrain vehicle operat-ing system, Some temporary simplifications were made to expedite operation of the coinpuler model along with changes to improve computer operation.
2.1.5.2 Controls Model
The logic incorporated into the model is essentially that described in the prior study,with the following exceptions;
• Engine characteristics are defined as algorithms rather than by partial derivatives to- provide better accuracy front the design points and is more efficient with respect to
computer time
• Variable combustor l eornetry is not included
• Only a simplified shift logic is presently incorporated in the program
• No active power split logic is included in the model, However, a passive power split.exists in which VSTC vanes are controlled as a function of temperature (either TIT orregenerator, whichever is closer to the limit)
• A major deviation from the prior study incorporates a clutch adjacent to the torqueconverter turbine. This was determined to be necessary for the differential gear systemoperating at idle and at very low vehicle speeds due to the high torque converte dragon the engine, The computer logic is for the clutch to slip at engine speeds less than54,5 percent and temperatures in excess of the set points (i.e., turbine inlet and regen-erator temperature set points)
2.1.6 Cycle Analysis
The RPD engine cycle analysis has been completed for maximum power and a vehiclecruise condition of approximately 64.4 km/h (40 mph) and is shown in Figures 32 and
t
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33, respectively, The cycle was established using a — 27 degree turbine exit swirl design,which provides the best drive<ability measured by the 4 second distance.
The engine provides a nominal 96,94 kw (130 horsepower) into the gearbox undermaximum power conditions obtained at 100,000 rpm turbine speed, Compressor massflow is 0,39 kg/sec (0,85 lb /sec) with a 5A pressure ratio, The turbine operates at1644K (2500'F) TIT w. th ;a regenerator inlet temperature of 1244K (1780'P'),
The cruise condition cycle provides approximately 11. 19 kw (15 hp) with a compressormass flow of 0,13 kg/sec (0,28 lb/sec) and a pressure ratio of 1,72. Regenerator inlettemperature limits of 1360K, (2000°F) cause the turbine to operate at 1516K (22701F)under this cruise condition,
RYD engine specific fuel consumption (sfc) characteristics are shown in Figure 34,Engine power up to 1.3,42 kw (18 horsepower) can be obtained, with a turbine idle speedof 55,000 rpna by modulating only the compressor inlet guide vanes (IGV), Power re-quirements above 0,42 kw (18 horsepower) require an increase in turbine speed and1()V modulation, A specific fuel consumption under 0.30 is obtained at power levelsbetween 27,59 and 82,03 kw (37 and 110 horsepower), Approximately 7,41 kw (10horsepower) can be obtained at a specific fuel consumption of 0,61 mg/'J (036 lbjhr-shp),
2,1.7 Vehicle Performance
Vehicle performance was established for the AGT engine using both the differentialVSTC and split-path VSTC gearboxes, Vehicle rear axle ratios were selected to providethe bast performance for each type of gearbox, A comparison of the CFAC mileage, 4-second distance, and 0 to 96,6 km /h (0 to 60 mph) acceleration time for cacti gearboxdesign is given in Table 3, Performance was established for a 1434,7 kg (3163 pounds)grass vehicle weight,
TABLE 3. AGT PERFORMANCE COMPARISON 1435 k9 (3113 LB) GVW
VEHICLE 0.96.6 km/HDESIGN CFDC 4-SECOND (0-60 MPH)
REAR SPEED RUEL ECON DISTANCE ACCELERATIONGEARBOX AXLE km/H km/LITER METERS TIME
TYPE RATIO (MPH) (MPG) (FEET) (SECONDS)
Differential 3.08 158,0 153 20,8 1119VSTC (98.2) (36.0) (6811)
Split-Path 2,73 157,6 153 20,0 11,6VSTC (97,9) (36,0) (68,8)
Fuel is gasoline with LHV of 43,838 kJ/kg (18,860 Btu / Ib),
46
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ss
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0.60.10
0,50.08
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0.04
1
15 150
kw
t 1 t t 1
2 10 20 100 200
SHP
'NET ENGINE OUTPUT POWER
Figure 34, SFC CharacterRstics — RPD Engine.
48
2. 1.8 Compressor definition
The compressor selected i'or further evaluation on the AGT engine is an aluminum,single-stage, centrifugal-type with a backward curved impeller, a radial vaned diffuserwith deswirl vanes, and variable inlet guide vanes (IGV), The radial IGVs are providedfor flow and pressure ratio control and power amplification at low speed off-design oper-ation, Mcridonal flowpath and station designation is shown in Figure 35.
Aluminum has been chosen for the impeller for low cost and ease of manufacturingconsiderations, An additional advantage of aluminum is the inherent lightweight and lowpolar moment of inertia, which aids engine response time.
Figure 36 presents "state-of-the-art" and 1983 projections for peak stage total -to-totalefficiency of machined and cast compressors operating at optimum specific speed. Figure37 reflects AiResearch experience concerning efficiency at speeds other than optimumspecific speed, "State-of«the-art" values are based on test results of several existing de-signs, including production units. The 1983 projection indicates that a 1.5 points effi-ciency improvement is expected to be achieved at a pressure ratio of 5:1.
It must be emphasized that performance levels in Figure 36 represent the maximum,attainable with no major design restrictions, In the normal series of compromises en-countered during development and the AGT "system design"; philosophy, peak compres-sor efficiency will most likely be sacrificed for the more important system goal; i.e., moreefficient low-power operation of the powertrain. In addition, operation at optimum spe-cific speed may be precluded by mechanical or impeller range considerations,
Cycle matching studies completed to date have produced a reference engine cycle forAGT engine system optimization. Aerodynamic and geometric features of the compres-sor corresponding to this reference engine are presented in Table 4.
This compressor has been selected witbt an objective of minimizing aerodynamic riskwhile providing performance that allows program goals to be achieved and/or exceeded.Important aerodynamic parameters, such as shroud relative diffusion ratio, inducer inletrelative Mach number, and radial diffuser inlet Mach number, have been set at levelsshown to result in successful compressors.
2.1.8.1 Compressor System Optimization
The AGT compressor system design was initiated by performing a parametric eudyusing the AiResea:rch centrifugal compressor preliminary design program to optimizecertain compressor design parameters. The preliminary AGT compressor mcridionalflowpath is shown in Figure 35. The following quantities were varied in the parametricanalysis:
49
F'
^ N lYut1 .
:l 1 n
4,2o0 — DESWIRL EXIT
3w890
DIFFUSER EXIT3.4
80
3.079
O AMBIENTu]2 , 6
-^i
A 60- — DIFFUSER INLET
2•.2 —IMPELLER EXIT50
1.88
4o-
1.4 SPLITTER LFADI:ZTO30 EDGE
01.0
200,6
1.0
-60-40 -20 0l l 1
20 40 60 T=L i 1 1
-2 -] 0 1 2 IN
Y 8
AXIAL DIMENSION
Figure 35, AGT Compressor Meridional Flowpath„
50
. P 8
iLi
O TEST DATA FOR EXISTING = 0.15CAST IMPELLERS MDIFF"EXIT
O TEST DATA FOR EXISTING C/b a 0.03MACHINED IKPELLERS
WV^1916m 1.68 . kg/s "(1.5 LB/SEC)0.85 \
OPTIMUM SPECIFIC SPEED
yy^^yU, \^^ \^\ AROil
SW 0.80u o ^ ^^ ^ O
0.75
2 3 4 5 6 7 8PRESSURE RATIO
c/b = CLEARANCE /BLADE HEIGHT
Figure 36. Current and Projected Compressor Efficiencies.
51
1/2N (QAV)
NS3/4'
(HALT)
rads/s (RPM)
m(FT)
(Q1 Q2) 1/2
w
w
kg/see (LB,/SEC)
kg/m 3 (LB/FT3)
10 20 30 40 50 60 70 80 90 100SPECIFIC SPEED- Ns (ENGLISH UNITS)
N 7..00U
4w
"t+
z
0.95HW44W
UH
0.90wHQ
NOTE: VALXD FOR 3<PRe.6
r 11
Figure 37. Adiabatic Efficiency Factor Versus Specified Speed.
52
TABLE 4
REFERENCE COMPRESSOR FEATURES
ImpNler
Physical Flow, kg / s (Ib / sec) 0.388 (0.855)
Total-Efficiency qt-t 0,897
Pressure Ratio, PT2 / PT 1 5.633
Temperature Rise, A T/ T O,703
Specific Speed, NS' 32.4 (58,6)
Number of Blades 12 Full/ 12 Splitters
Slip Factor 0.949
Compressor Inlet Temperature, 302.6 (544.7)
Stage
0.805
PT3 / PT 1 5.0
T1,K (OR)
Compressor Inlet Pressure,P1, Kilopascal (PSIA)
SPA, NPHYS (RPM)
Tip Diameter, mm (in.)
Inducer Hub Diameter, mm (in.)
Inducer Tip Diameter, mm (in.)
Inducer Tip Normal Thickness,mm (in.)
Exit Blade Angle, Degrees
Axial and Radial Clearances,mm (in.)
'NS is Defined as:
N G 3/4 WHERE:HACT
99.81 (14.476)
100,000
108.229 (4.261)
30.5 (1,2)
65.735 (2.588)
0.38 (0.015)
50
0.076 (0.003)
N = RPM
QAVG = (Q 142) 1 / 2 , m3/s (FT3 / SEC)
HACT = HEAD, m (FT)
53
• Impeller blade exit angle (backward curvature)• Shroud exit/inlet relative velocity ratio• Vaneless diffuser exit Mach number• Radial diffuser exit Mach number• Impeller blade count
The following quantities were held constant for the analysis:
• lnlet shroud/mean nieridional velocity ratio (01mrat) of 1.25• Hub radius 15,24 min (0,600 inch)• Deswirl vane exit angle (0 degrees)
Air impeller with splitter blades was used for this study,
Analytical results, for a radial diffuser vane exit Mach number of 0.25, an impellerhaving 12 full blades and 12 splitters, and a diffuser inlet Mach number of 0.9, areshown in Figure 38. Shroud relative velocity ratio was varied from 0,5 to 0.6 withvirtually no diffetcoce in performance. The strong effect is the rotor exit blade angle,which indicates continued efficiency improvement up to the highest angle (50 degrees)for which calculations were made,
Parameters for the optimization of the diffuser system, which has not yet been com-pleted, are shown in Figures 39 and 40, An impeller relative velocity ratio of 0,6, radialdiffuser vane exit Mach number of 0,25, and vaneless space radius ratio of 1,075 wereheld fixed, In all cases, the deswirl vane exit angle (Figure 35) was set to 0 degrees whilethe deswirl vane two theta (20) diffusion rate was varied from 6 to 12 degrees.
The analysis again shows peak ef'f'iciency to occur for the maximum impeller exitblade angle, A. two theta (20) of 6 degrees gives the best stage efficiency, Further analy-sis is to be conducted using a revised deswirl vane exit loss model which accounts forturning of the flow,
Blade count study results are shown in Figure 41, The exit blade angle for this calcu-lation was 50 degrees, diffuser exit Mach number was 0.2, and inlet Mach number washeld at 0.85. The trend indicates increasing efficiency for decreasing blade count. Final.selection of blade count is discussed in the following paragraphs,
2.1.8.2 Initial Impeller Blade Design
The following parameters were selected to b ,^;gin initial impeller blade design:
• Exit blade angle, 40 degreesShroud relative velocity ratio, 0.6
N Y
• 12 full blades, 12 splitters
54
0.90IMPELLER
0.85
STAGE
W2/W10.500.55 --^—0.60
0,8b
RADIAL DIFFUSER INLET MACH NO = 0.90
RADIAL DIFFUSER EXIT MACH NO = 0.25
IMPELLER: 12 FULL, 12 SPLITTER BLADES
VAT = 1.25
0.75 L30
40 50
60
IMPELLER EXIT BLADE ANGL'^C
Figure 38. Efficiency Versus Impeller Exit Blade Angle.
1
HH
UZwHUHwww
55
0.82iH0.81
tiW
v 0.80cn H
w 0.79W ^
0..20
Wz0.15
H3z
A 0. 10
-tzai^DESWIRL VANE EXIT ANGLE a 0 DEGREESDESWIRL VANE ASPECT RATIO = 1W2/W1 n 0.6
DESWIRL VANE THICKNESS a 0.5 mm (0.020 IN )DESWIRL INLET MACH NO a 0.25VANELESS SPACE RATIO a 1.075
o a.0
1.5cn
41.0
DESWIRL 2 9
6 -8 __•_.._10------12
0w
a 607 LO 185Da e 590 N 1800
H H574 w 1750
0.6
Fa O
H u 0.5
o q 0.430 40 50
IMPELLER EXIT BLADE ANGLE
Figure 39. AGT Compressor Optimization.
56
i
EH 0.82t^
0.81
w DESWIRL VANE EXIT ANGLE = 0 DEGREESc^ H 0.80 DESWIRL VANE ASPECT RATIO = 1OC
H W2/W1 = 0.6w 0.79
W DESWIRL VANE THICKNESS = 0.5 mm ( 0.020 IN )DESWIRL INLET MACH NO. : 0.25VANELESS SPACE RATIO - 1.075
w
IN ,MM
H 4.0 102
3.5 s -.►^'^^^.,.^'' DESWIRL 20
x8 ---- --
o w 3.0 7 10-- --12
50
Z
Rp►pIApIF^t3iEi i
H
Z w 40'HaW 30
p DESWIRL
20 VANE
aw ^,'^ z 1.0w^A 0.9aN aW 0.8
H 30 40 50
IMPELLER EXIT BLADE ANGLE
Figure 40. AGT Compressor Optimization.
57
E:
E-0.86
HU►-i 0.84wwW
0.88
0.82
0.8020
0.90
22 24 26 28 30 32
IMPELLER EXIT BIADE COUNT
Figure 41. Efficiency Vorsus Impeller Exit Blade Count for AGTCompressee Design.
58
An exit blade angle of 40 degrees was the limit set by the assumed maximum allow-able blade stress-, aerodynamically, a higher angle would be desirable, Selection of ashroud relative velocity ratio of 0,6 was based on past experience which indicates betterrotor discharge velocity profiles when iess diffusion (higher relative velocity ratios) oc-curs in the impeller, Impeller exit blade count was set at 24 exit blades, since a loweroumlaer did not offer significant gains in stage efficiency, Final decision on blade countwill be made based on blade surface Mach No, distribution,
Two inlet flowpaths were evaluated during this analysis as shown in Figure 42, Theshortened inlet was required to improve engine rotor dynamics as discussed in Section:?.1,2,3, The effect on compressor inlet surface velocities is compared in figure 43, Theshort inlet effectively eliminated shroud diffusion and reduced hub diffusion, which aredetrimental to the surface boundary layer,
Initial blade geometry design indicated a high blade stress at the hub trailing edge,Therefore, blade root thickness was increased in this region from 0,635 to 1,016 mm(0,025 to 0,040 inch) and a new blade generated, Resulting blade stresses are shown inFigure 44.
Blade surface and meanline velocity distributions are shown in Figures 45, 46, and 47,Along the shroud streamline, efforts were made to reduce the diffusion rate near theleading edge and, thereby, limit overshoot in suction surface velocity, This helped toassure adequate impeller surge margin, On the hub streamline, pressure surface diffu-sion from the leading edge to the minimum value, which occurs just ahead of the splitter,was minimized, This is advantageous for impeller surge margin and efficiency considera-tions. These adjustments were made by changes in area control oaring parameters suchas blade shape, blade angle and flowpath geometry,
2.1.8.3 Compressor Test Rig
The compressor test rig layout is shown in Figure 48, Power is derived from a driveturbine system composed primarily of the AiResearch Model 'I"V81 Turbocharger hotflowpath components. Actual power section compressor stage hardware (VIGVs, impel-ler, and diffuser) is utilized in the rig to permit detailed compressor design performancemapping, As shown in Figure 48, power section ducting has been incorporated to estab-lish any effects of flow phenomena resulting from transitioning compressor dischargeflow to the regenerator high pressure inlet, In addition, the inlet air filter and associatedducting (not shown) can be accommodated to verify aerodynamic design intent andagain establish any detrimental flow effects resulting from the installation,
Compressor rig rotor dynamics have been analyzed and are presented in Figures 49and 50. Using bearing stiffness values of approximately 4378 kN/m (25,000 lb,/in.), thefirst three undamped critical speeds are 27,000, 49,000, and 180,000 rpm, respectively.The first and second critical speeds are predominantly rigid body modes, whereas the
59
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OC,4
0
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Z L I 1 f I MwMMMMMMj
H00
r[ c;
SIIICIVH
Figure 42. AGT Compressor.
60
FT/SEC m/s6QO r 180
ROUD
Soo r 150
400 1
120
HLF-i
.z
300 90
r_H
F^(
200 r 60
l
1.00 30
0 1 C0
UB
20 40 60 80 100
PERCENT MERIDIONAL DISTANCE
Figure 43. AGT Compressor Inlet Surface Velocity
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third critical is a bending mode, Past experience with hydraulically mounted bearingsystems has indicated that the first two criticals will be sufficiently damped and poselittle problem during operation of the test i;, The third critical has sufficient marginabove the operating envelope and is consistent with AiResearch Phoenix design technolo-
gy,
The rig incorporates the capability for controlling and evaluating critical compressorrunning clearances based on. capacitance probe measurements at selected stations alongthe impeller rneridional flowpath. The clearance control mechanism is based on a designconcept successfully used in the improved surge margin centrifugal compressor programfor the U,S, Army Air Mobility R&D Laboratory, This mechanism features an ACMEthread, used to aaiaty adjust, the position of the compressor rotor, This is accomplishedby fixing the internal thread and turning the external thread, thus forcing the externallythreaded members, which are connected to the impeller, to move axially, A cylindricalworst gear set is used to drive the external ACME thread, The splined drive shaft, forthe compressor is required to move axially relative to the external ACME thread due toaxial positioning of the compressor, deflections, and thermal gradients, Friction in thesplines creates high forces that resist this movement and induce high loads on the bear-ings, worm gear, and ACME thread, Therefore, a ball spline is incorporated to minimizethis axial load from spline friction,
Clearance control capability will allow thermal stabilization of the test rig at test pointconditions, followed by a reduction of running clearances, to the small desired values10,076 min (0,003 inch)]. This mode of operation will allow testing to proceed in a safe,rapid, and cost-effective sequence,
The compressor rig incorporates extensive instrumentation to allow performance eval-uation of all stage components as well as overall parameters, This will include,
• Inlet total pressures and temperatures• Shroud line static pressures• Impeller exit static pressures• Vaneless space static pressures• Diffuser static pressures• Stage exit total pressures and temperatures• Inlet. and exit flow measurement.• Provisions for stage exit surveys• Clearance probe measurements
instrumentation is of a size consistent with the small stage dimensions and of suffi(i,nt number and spacing to provide realistic average values for all parameters mea-sured,
2.1.9 Regenerator Definition
66
}
6at
1y M
{ 8
ORIGINAL PAGE ISOF POOR QUALITY,
Figure 48. Compressor Test Rig Layout.
7 7
r
67
150H
tiaa
viw
P4U)
a
HHP+U
cR^^^^U
100 PERCENT'OPERATING SPEED
10 --
5
20
Y
3^d CRS21CA
01M
0 1-- I ► _i
0 0.5 1.0 1.5 2.0 2.5 3.0 3.5 kN/m x 10^^
t l I i _
- 1
—40 5 10 15 20 LB/IN x 10
BEARING STIFFNESS
Figure 49. Compressor Rig Critical Speeds.
68
.ro IN
MM
MODE 1
M,OY ^.m a.U3 f.uu SAW 6.00 1.00 6.06 11.00 10m IN
50 100 150 200 250
MM
AXIAL DISTANCE
AXIAL DISTANCE
MODE 3
1'.ao z;m a;oo uo ti;m c;m coo f,m fo.
50 250100 150 200
IN
mm
NHHHau
°za
woNHHHa
^Wowz
N HH E+aUw
ZA
MM IN
25.4 0.50-
0.0 0.00•
-12.7 -0.50
25.4 1.00
0.50
0.0 0.00,
-0.50
-25.4--l-00
25.4 1.00
0.50
0.0 0.00
-12.7 -0.50
AXIAL DISTANCE
Figure 50. Compressor Rig Rotor Dynamics.
69
2.1,9.1 Regenerator Core and Seal Configuration
The initial regenerator size selected for the design study was 444,5 Inn U p (17,5 in)x 139,7 tuna ID (5,5 in) x 71,12 nim (2,8 in) thick. A performance objective was estab-lished yielding 919 percent effeetivcness with 7,5 percent pressure drop at maximumpower conditions.
This performance is based on an extruded isosceles triangular matrix configurationwith .a 4,572 nine (0,018 in) hydraulic diameter (IM) and 0.0889 mill (0,0035 in) ma-terial thickness that is expected to be developed by 1983, At the present tinge a sinusoi,dal triangular structure fabricated by horning g lass with 0,001 nim (0,0024 in) materialwell thickness represents the best state-of-the-art fin configuration for the ACCT engine,For the selected core size an effectiveness of 94,2 percent with 6,25 percent pressuredrop can be expected at fell power conditions, Since the pressure drop for the state-of-the-art fin geometry is considerably less than the design objective, the regenerator thick-ness can be increased, By increasing the thickness from 71, I? anall (2,8 in) to 83,82 mill(3,3 in) the state-of-the-art matrix should increase regenerator +affect veness from 90,2 to91.9 perutlt with a pressure drop of 7,3 percent \which is still below the design objectiveof 7.5 percent, By increasing the hydraulic diameter requirement for the future extrudedisosceles triangular stmeture from 0,457 mall (0,018 in) to 0,508 nim (0,020 in) for thethicker regenerator size, the original performance objective can be maintained, Based onthese results the regenerator sire recommended 1'or continued study was 444,5 mm OD(17,5 in) x 139,7 mara 11) 1,5,5 in) x 8:3,82 nim (3,3 in) thick,
A
During the study a regenerator performance analysis was conducted to evaluate theeffect of flow area unbalance oil effectiveness arad pressure drop, This pre-liminary hiformation was required in order to determine the initial regenerator inner andouter seat crossarnl shape, Three inner seat design configi ations were evaluated as illus-trated in Figure 51. Configuration No, I is similar to the seal developed for the FordStirling engine, The equal flow area split (balanced heat exchanger) provides the Bestregenerator effectiveness and pressure drop combination throughout the engine operatingrange, Although configurations No, 2 and 3 offer certain advantages from the standpointof seat fabrication and leakage, a significant increase in regenerator core diameter wouldbe required to attain equivalent effectiveness and pressure drop with respect to configur-ation No. 1, For this reason, configuration No, 1. was selected for continued design eval-uation,
Further system optimization studies regarding secondary cooling air for the combus-tor, indicated an increase of regenerator core inner diameter was required, The innerdiameter was increased from 139,7 nanl (5,5 in) to 160,48 nlm (6,318 in), To maintainthe same regenerator performance the outer diameter was increased to attain equivalentfrontal area. As a result the regenerator recommenfled for the AGT RYD engine is454,05 nim OD (17,876 in) x 160.48 nlnl ID (6,318 in) x 83,82 mm (3,3 in) thick,
70
CONFIGURATION 1
CONFIGURATION 2
CONFIGURATION 3
Figure 51. Regenerator Inner Seal Configurations.
71
As an alternative to the extruded isosceles triangular matrix with a hydraulic diameterof 0.508 Warn (0,020 in), a design study was completed for optimization of an extrudedrectangular matrix with a hydraulic diameter 0,5134 mm (0.023 in) to 0.610 mm (0,024in). NO designs should attain equivalent performance, This study also established thetooling requirements for the regenerator core suppliers, Meetings were held with repre-sentatives from Corning t,lass Works and NGK, Insulators to discuss the status of theirmatrix extrusion process. Both companies expressed confidence in developing this processto produce the thin-wall regeneratoe matrix structures, which are required for compacthigh performance heat exchangers,
A two-dimensional heat transfer analysis was initiated to define the temperature andthermal stress distribution for the regenerator assembly,
2.1..9.2 Regenerator Drive System
Based on prior lord experience, 'fable 5 contains generalized gear geometry that hasbeen established for the regenerator pinion and ring gear,.
TABLE 5.. REGENERATOR PINION AND RING GEAR DATA
I
TeethCenter Dist, mm (in.)Diameter PitchPressure Arogle, DegreesWorking Add, mm (in.)Dedendum, mm (gin.)Tooth Thick, mm (in,)
Tip Thick, mm On,)
Face Width, mm (in.)
Pitch Diameter, mm (in.)
Bonding Stress, KPA (psi)
PVT Factor, nm / sec / cm (Ib ft/sec/in.)
PV Factor, nm / sec (lb ft/sec)
Peak Torque, nm (lb ft)Peak Speed, (rpm)
Pinion
30
1822,5
1.35 (0.053)1.78 (0.070)
2,35 (0.0924)
1.12 (0,044)
42,3 (1.667)
68223 (9896)
564,2 (1638)
18540 (13673)
23.86 (17,6)
284
Gear
341
261,8 (10,306)
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2.36 (0.093)
1,85 (0,0729)
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25,4(l.00)
481.2 (18.944)
61287 (8890)
436.0 (1266)
18920 (13953)
271.2 (200)
25
72
The regenerator core torque requirement at full power engine conditions was estimat-ed based on 14 previous analysis performed on the Ford 707 engine regenerator system,Since the seal configuration for the AGT engine has not been finalized, this approachrepresents ;an initial approximation, The 707 engine analysis was modified based on theAGE' engine pressure and seal width and radius,
Utilizing a design core torque requirement of 200 lb-ft at full power condition thelocation of the fixed and spring roller was optimized with respect to the present pinionlocation and inner and outer seal crossarm orientations, The design criterion selectedwere as fellows;
1, Fixed roller reaction load to be less than half of the 707 engine, which incorporates ayoke to accommodate the total load,
2, Minimum spring roller load to ensure ring gear engagement. with the pinion,
3, Minimum force transmitted to the fixed roller due to increased spring roller loads,
By minimizing the fixed roller loads, the graphite bearing system that has been suc-cessful in the 707 engine can be incorporated,
Subsequent analysis was completed to determine the roller reaction loads resultingfrom a variation in torque requirements of 50 to 400 lb-ft, The results illustrated onFigure 52 assume ;Aiding friction at the pinion roller location and rolling friction at thespring and fixed roller locations,
2.1.9.3 Regenerator Test Rigs
Preliminary layout design of the two regenerator cold flow rigs, shown in Figures 53and 54, has been completed, The objective of these rigs is to obtain a now match be-tween the regenerator cote high and low pressure sides through various header designs(Figures 55, 56, and 57), Complete velocity distribution mapping immediately downstream of the core will be accomplished through use of hot wire anemometers in which auniform radial and circumferential velocity profile is desired.
To obtain the desired flow match, an iterative test approach will be utilized. Tangen-tial and radial flow patterns resulting from various header designs will be used as inputinto a 3-dimensiona1 nodal analysis to obtain immediate feedback on test configurations.Therefore, optimum design can be derived with effects on core and engine performancepredicted throughout the engine operating range.
2.1.10 Combustor Definition
Definition of a single-can, variable-geometry, ceramic combustor was initiated, Primecombustor design was derived from Ford variable-geometry combustor technology and
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Figure 57. Air Side Alternate Header Design.
79
AiResearch empirical/analytical oombustor design approaches, As a goal, the combus-tion system will be below the following emission levels for a fuel economy goal of 15.3km/t" (36 mpg) using gasoline:
Oxides o1' nitrogen, NOx
Total parrticulates
0,25 gm/km (0,40 gin/mile) orless than 5,7 gm/kg fuel
0,26 grn/krn (0,41 gm/mile) orless than 5,13 gm/kg fuel
2.11. gm/km (3.4 gin/mile) or44,8 gm/kg fuel
0.12 gm/kni (0.2 gan/mile) or2.6 gm/kg fuel
Total unburned hydro-carbons, HC
Carbon monoxide, CO
In addition to emission goals, the combustor will have the ability to use a variety of:alternate fuels. The following combustor performance parameters Lave been establishedto assure that the selected combustor is compatible with the selected engine concept.
Combustion efficiency 99 percent at all engine operating conditions includingaltitude cold start, acceleration and deceleration
• Combustor pressure drop 3 percent
• Combustor exit patternfactor 0.15 at the design condition
Based on Ford work with several different combustor riga in 1973 and 1974, it Wasestablished that low ernissions could be attained throughout the operating range of theFord Model 820 turbine engine by using a premixing, prevaporizing combustion systemwith a properly designed variable-geometry actuation system. The Ford design is shownschematically in Figure 58. Combustor flow conditions for the IGT engine were suffi-ciently close to the Ford Model 820 turbine engine to afford scaling. Combustor scalingresults are shown in Table 6.
80
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81
TABLE 6
COMPARISON BETWEEN FORD VG COMBUSTOR ANDPROPOSED AGT COMBUSTOR
1
FORD PROPOSED
COMBUSTOR COMBUSTOR COMMENTS
158.75 101.6(6.25) (4.0)
215,9 146,05(8.50) (5,75)1,36 1,44 PF
1.83 0.43 Idle Emissions(0:114) (0,027)0,63 0. 40 Idle Emissions
(0.039) (0.025)7.1 4,7 NOx
888.9 1344.4 Idle Emissions(1600) (2420)
Combustor Diameter,mm (in,)
Combustor Lengthmm (in.)
Combustor L / D
Loading Parameter atIdle, kg! s-m3(PPS / FT3)Design, kg / s-m3(PPS / FT3)
Residence Time, MS
Idle TIT, K (OR)
In describing the combustor, it is convenient to separate its operation into three are.;.s;prevaporized - pre aixer channel, variable-g r oi .ietry feature, and fuel nozzle, The follow-ing discussion describes various phases of design methodology and presents the latestcombustor configuration.
The purpose of the nrevaporized-premixer channel is to vaporize liquid fuel prior toentering the burning zone, A problem encountered in the vaporizing zone is the potentialof flashback or premature burning in the prevaporizing channel, ] lord successfully timi-nated this problem through selection of an appropriate reference velocity in the premix-ing zone. Thus, the AGT reference velocity of 31,21 m/sec (102,4 ft/sec) andprevaporizing diameter of 58.4 mm (23 inch) were scaled from the Ford configuration.The prevaporizer length was established by calculating the distance required for the fueldroplets to completely vaporize without impinging on a solid boundary. Two spray coneangles (45 and 70 degree) were studied in conjunction with two initial droplet sizes;Sauter Mean Diameters (SMD), of 50 and 25 microns. Results of this study indicatedthat a prevaporizing length of 58,4 mm (2.3 inch) provides adequate distance for thefuel droplets to vaporize at all operating conditions, including a margin allowing for thepresence of excessively large diameter droplets. The importance of complete vaporizationprior to wall impingement can be seen from information based on pre-ignition studies oninternal combustion engines indicating that ignition will occur if fivA contacts surfaceswith temperatures in excess of 11440 (16000F).
82
The importance of the variable geometry mechanism for the AGT combustor is para-mount, (changes in geometry should duplicate the area changes required, as well as theability to respond quickly, Since the AGT primary zone fuel-air ratio will be lean forNOx control, lack of a quick acting variable geometry system could either increase thefuel-air ratio, which would cause an increase in emissions, or decrease the fuel-air ratiocausing blow out,
Prevaporizer and secondary orifice flow split is dictated by emissions considerations,The control of unburned hydrocarbons (HC) and carbon monoxide (CO) can be accom-plished by assuring uniform and complete mixing of the fuel and air, This should beaccomplished by proper injector-prevaporizing zone design, The formation of nitrogenoxides (NOx) can be minimized by burning at very lean fuel air ratios. Work presentedin NASA conference publication 2078, Figure 59, reported that if the maximum com-bustor temperature is kept below 1922K (3000°F) for a residence time not exceedingtwo milliseconds, the amount of NOx produced is equal to 2,0 kg NOx/ 1000 kg fuelassuring achievement of the program goals, To maintain a 1922K (3000°F) combustortemperature, the flow-split between the prevaporizer and secondary orifices would be asshown in Figure 60,
The combustor burner metering open area will not only be varied to obtain the desiredflow-splits, but also to maintain a combustor pressure drop of three percent of combustorinlet, pressure, Since inlet conditions to the combustor vary considerably, depending onthe power setting, some compromises will be necessary, Combustor inlet conditions ver-sus shaft horsepower are shown in Figures 61 through 65. Variations result frorn differ-ent compressor VIGV settings in correlation to different engine controlling limits depen-dent on speed. Thus variations in inlet conditions will necessitate that nominal designemissions goals be set sufficiently below program goals to ensure that emissions encoun-tered at the more severe operating conditions will not exceed overall driving cycle goals,
Present baseline combustor configurations are shown in Figure 66, The fuel injector insubjected to combustor pressure drops under all conditions, and primary zone airflow isincreased by metering flow through a swirler concentric about the fuel nozzle, Swirlingprimary air should increase the rate at which air and fuel mix thereby creating a morehomogeneous mixture. The main burner for both combustor configurations is identicalexcept for differences in the pre7nixer-vaporizer section. One configuration employs anaxial swirler vane surrounding the fuel nozzle, whereas the other design uses a radialinflow swirler surrounding the fuel nozzle. The radial inflow swirler configurationreduces th;; fuel nozzle length immersed in hot combustor inlet gases as well as simplify-ing variable geometry controlling devices.
The swirler metering section, center body plunger, and dilution zone metering are alllinked together. As fuel flow increases, the plunger moves in, a,,ti does the swirler meter-ing block, increasing air flow to the swirler. At the same time, the secondary area isclosed off mai> ta'ni,ig approximately the same total combustor open area. Methods forcalculating both van rizer and co , -bustor size were accomplished, In the first step the
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40
55
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0 20 40 60 80 1,00 120 140 160 180 kW
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Figure 61. Effect of Engine Power on Combustor Fuel Flow.
86
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Figure 62.. Effect of Engine Power on Combustor Airflow
87
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88
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91
combustor effective open area for various metering locations was calculated for h g dif-ferent sets of combustor inlet conditions, hosed on predictions by the performance model.Points covered were front to maximum lower, including acceleration points, Variousareas were calculated to yield it primary zone temperature of 1922K (3000'F), Plots ofvarious open areas its a function of fuel flow tire shown in Figures 67, 68, and 69. As Canbe seen, area variation for it fuel flow is significant necessitating fitting a curvethrough the area irauts such that one distinct area existed for a given fuel flow as shownin figure 70, With this area variation, the model was then used to predict the variationin primary zone temperature, shown in Figurc 71, Points above 2177K (3460'F) coin-tided with short time period (approximately 2 seconds) engine acceleration points, whichshould have little effect on the overall cycle emissions,
AU`T' fu0 nozzle development offers it unique challenge, The high combustor inlettemperature environment could cause fuel eegradation and cracking within tilt: nozzleresulting in plugging and reduced auto ignition effectiveness, Also, pre-ignition can occurif the fuel-air mixture contacts any surface ex.cceding 1144K (16000F),
The preliminary fuel injection system design under investigation for the AGT combus-tor differs from the I ,ord system in that it single, centrally located fuel nozzle is Used,whereas the Ford system employed three individual nozzles, Prclianimtty studies haveshown that the single: nozzle located concentric about the shaft will produce it moreuniform eircumferwitial fuel distribution, A conceptual design is shown in Figure 72,(.cycle tmalysis is currently being conducted regarding effects of compressor bleed its asource for air assist,
Surrounding the airblast fuel injector is a swirler with turning vanes set at 30 degreeswith respect to the combustor axis. The swirling airflow .around the nozzle enhancesmiring of the fuel and ,air creating it uniform fuel-air mixture, thereby minimizing localrich and lean regions, The tentative flow split within tfae nozzle is as follows;
Percent of TotalPassage Nozzle Airflow, Percent
inner airblast 23,9Outerairblast 37.0Shroud 39A
The ratio of inner , jyoutcr airblast airflow wits selected based on a survey of flowsplitters in various airblast nozzles currently being used at AiResearch Phoenix, Shroud.air was selected to be approximately two-thirds of the inner and outer airblast air cont-bined. This should produce it spray cone angle between 50 .70 degrees as demonstratedon a previous combustor program,
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To avoid preignition, it will be. necessary to keep vaporizer residence time below theignition delay tune, As showy in Figure 73 ignition delay for octane is 0.7 millisecondsat 1366K (2000 0 F), Uator isken at 1477K (2200°F) indicated a delay time of 0.2 milti4
seconds, Short vaporizer residence time impact on the fuel injector is such that the injec-tot mus± supply it finely atomized spray on the order of 20 micron SMD to achievevaporization of the majority of the fuel in the 0.7 millisecond range. The effect of initialdroplet size cn time to evaporate the fuel is shown in Figure 74, For a 0,7 millisecondresidence time, mamnium allowable fuel droplet size is 24 microns, which would iiecessi-tate an average droplet size of approximately 15 microns, At the present: time, severalnovel fuel injection devices are being evaluated at NASA-URC that show promise in avaporizer application, However, airblast injection is a more proven method of injectingfuel at a small droplet size over a wide range of fuel flows,
Due to design complexity associated with the fuel nozzle, the decision was made tocontact nozzle vendors for design concepts, A nozzle envelope, similar to a previouslymentioned AiResearch design, and nozzle inlet conditions at various engine conditionshave been given to several fuel nozzle vendors, Specific nozzle requirements include goodatomization at all operating conditions including starting, minimal fuel pressure dropacross the nozzle, and the possibility of internal nozzle cooling including both air andfuel azs the cooling media.
2.1.10.1 Combustor Test Rig
A test rig definition has been completed to facilitate testing of the variable ge)metry
Itcombustor, The test rig will be built in two configurations with relatively minor raodifi-cations required to convert from one configuration to the other,
Figure 75 shows a conceptual drawing of the first build, Thermocouple probes are notshown, but it is intended that these be located at a point near the burner discharge, Thepurpose of this build is to verify combustor operation and establish. a temperature pat-tern at the burner discharge,
Figure 76 is a conceptual design for the second build, The duct section of the firstbuild has been replaced by a nose cone, the purpose of this configuration is to simulatethe engine flow path prior to the turbine and acquire temperatures and pressures at apoint that will coincide with the turbine stator.
2.1.11 Turbine Definition
2.1.11.1 Preliminary Investigation
Need for high overall performance over the wide range of operating conditions for anautomotive application requires that component performance be high at the idle andintermediate power operating conditions where the most fuel is consumed. However, to
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Figure 74. Effect of initial Droplet Diameter onthe Evaporation Time.
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Figure 76. ACTT Combustor Toot Rig.
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meet vehicle driveabilaty c'-.aracteristics, the turbine must be capable of producing maxi_muna power at the c cle temperature limit and 100-percent shaft speed, This require"anent leads to choosing a designs point between the maximum power and minimum sfeoperating, conditions, This is in marked contrast to most gas turbine applicaations whichrequire maximum efficiency at maximum power conditions and produce lesser efficien-cies part-power performance.
To investigate the effect of turbine off-design operating conditions on engine perform.once, three design point vector diagrams were eefined at the maximum power condition,Tile vector diagrams are identified by the degree of rotor absolute exit. <iwirl (i,c, —9.8, 27 reference and 37 degrees). Efficiency and exit swirl characteristics front engine idleto naaxinau ►» ficiwcr are fresented in Figure 77 based on the turbine off-design computerprogram, As engine: sped is reduced for Bart power operation, the amount of turbineexit swirl increases (becomes more positive j. The lowest swirl design has the highesteffieieaa^y at, full speed, ', o, (he lowest efficiency at low speed, The highestr negative swirl.angle design bas tine lo>,:st efficiency at .maximum power and the highest efficiency atidle conditions, The reference design turbine is a good compromise, and has been thehaask of all e,qt lne performance presented to date. lJowever, higher efficiency at idleconditions, available from a high swirl (-37 degrees) turbine, is a potential advantage todriveability and fuel ecoraowy, 'T'lae reference swirl turbine can be derived froara, the highswirl design by a simple cutback of the exducer blades,
I)ef, nition of (lies turbine stage has been concentrated in the following Fareas, statorinlet l'lowpath, stator els.sil;aa, rotor design, and rotor stress anaalysis, The paragraphs thatfollow discuss these efforts.
2.1,11.2 Stator Design
2.1 ,1 '1.2,1 Stator Inlet F'lowpath
stator inlet duct definition consisted of examining alternate hub and shroud con-tours, axial lengths, and combustor exit transitioning, "fable 7 surnrnarizes geometries Athrough 0 examined to date and lists rating criteria used for the study (i.e,, perk veloc-ity which occurs on the stator hub inlet bend and the; velocity gradient across the statorinlet, plane), Configuration A was run with an increased axial length of 12,7mm (0.50inch) to evaluate; the inipact of envelope constraints. The velocity distribution shown inTable 6 was not significantly improved, compared to the impact of adjusting the engineconfiguration, Therefore, an increase in axial length was not considered necessary toarrive Fat, a sFatisf'tactory inlet duct velocity distribution,
C'onf'igurations B and C examined the tradeoff between peak velocity ( ►maximum cur-vature) along the hub to tlae stator inlet velocity gradients, If the hub fiend radii isdecreased (Configuration l3), tlae stator inlet contour can be rtiaintained radial (90 de-gree inlet). This results in a higher velocity peak and diffusion, but a more uniform
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PEAKVELOCITY
m/s(FT/ SEC)
7912(260.0)
STATOR INLETVELOCITYGRADIENT
m/s (FT/SEC)
29.6(97.0)
stater inlet cross passage velocity profile, Configuration B resulted in a peak velocity of1033 m/s (339 ft/sec), reducing the stator inlet velocity gradient to 384 m/s (126.0ft /set` ),
TABLE 7. STATOR INLET FLOWPATH CONFIGURATION STUDY
CONFIGURATION DESCRIPTION
A Duct axial length(combustor exit tostator) increased by12,7 mm (0,6 inch)
B Original axial length,slope at stator inlethub and shroud contour
C Original axial length,90 degree stator , Inlethub and shroud contour
D Original axial length,stator inlet configur-ation between B and C,Blend at combustorexit shroud region,
90.8
46,9
(298,0)
(154.0)
103,4
38,4
(339.1)
(126,0)
W3 42,7
(293.0)
(140,0)
The other approach is to n`aaximize the hub bend radius with a wall slope at the statorinlet (Configuration C), Table 6 shows that peak velocity is reduced to 90,$ m/$ (298ft/sec) and the stator inlet gradient. increases to 46.9 rn/s (154 1`t; jsec), Configuration Dis a compromise between Configurations B and C and incorporates combustor exit-ductinlet blending in the shroud region,: Figure 78 shows this configuration and also thelocation of six centerbody struts (NACA 16-021 profile) incorporated in the flow Solu-tion model, The velocity distribution with these features is presented in Figure 79, Thenext phase of the program will continue examining; contour modifications and evaluatingthe hub and shroud boundary layer characteristics with two-dimensional methods, Thegoal is a non-separating design with minimum shroud diffusion and rninintum suatorinlet, velocity gradient,
L.
106
^m t^' SI
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100
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as8 —STATOR
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U7
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CWPE12lODY
ASSUMED STAGNATION CENTERLTNE
0.0 Lo.o0 25 50 75 mm
t01.0
i1.0
i2.0 3.0 IN
AXIAL LENGTH
Figure 7S, 'Turbine Inlet Duct, Configuration D.
107
. 5 HUB STREAMLINE
USTOR EXIT PLANE
25 50 75 100 125 150 11411
I- 1 1 I 1 i I
6.0 IN0 1.0 2.0 3.0 4.0 5.0
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M, ME'RIDIONAL DISTANCE
Figure 79. Turbine Inlet Duct Velocity Distribution ForConfiguration D.
108
Vu/.Acr - 0.7660
[1--Q
'Acr - 0.136^- STATOR INLET,
1 72.67° R - 90.9523 mm(3.5808 IN )
/Acr - 0.2389 Acr - 730.62 m/s/-^ (2397.06 FT/SEC)
- STATOR EXIT,R - 70.9855 mm
(2.7947 IN )Acr - 730.62 m/s
(2397.06 FT /SEC,)72 DEGREESd 02- ROTOR INLET,Q2 =I -29.16 DEGREES R - 66.9442 mm
11(2.6356 IN )W
-/Acr - 0.092 (0.302) Ut 701 m/s S
1 Acr 730.62 m/s T 3
- ( 1644K (2500 0 F) (2397.05 FT/SEC) IN----p3 - -64.38 DEGREES
a„ - -37.06 DEGREES
%
cx fAcr - 0.0804IN
R (0.2639)u/Acr - 0.8122U/Acr - 0.9595
Vx/Acr = 0.1291(0.4236)
O- ROTOR EXIT MEAN,R = 34.2748 mm
(1.3494 IN )Acr = 686.76 m/s
(2089.12 FT/SEC)
C$ h'r0<<s^ o
u/Acr o^^^,0.3 2 0 'j
U/Acr 0.5637Wu/Acr - -0.8836
V = ABSOLUTE VELOCITY, m/S(FT/SEC)
Vu ABSOLUTE TANGENTIAL VELOCITY,m/S (FT/SEC)
U = ROTOR WHEEL SPEED, m/S(FT/SEC)
W RELATIVE VELOCITY, m/S(FT/SEC)
Wu RELATIVE TANGENTIAL VELOCITYm/s (FT/SEC)ABSOLUTEFLOW ANGLV, DEGREESRELATIVE FLOW AN=4f DEGREES
NOTE: THE STATOR INLET RADIUSAND VELOCITY ARE SHOWNWITH A PARALLEL END WALLAND 17 VANES
Figure 80. Turbine One-Dimensional Vector Diagram ForHigh Swirl (-37.0 Degree) Design Point.
109
Figure 81, Preliminary Geometry For 18 Vane CeramicRadial Turbine Stator.
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2.1,11.2.2 Stator Aerodynamic Design
The stator and rotor detail design described in the following sections is based on thehigh exit. swirl (-37.0 degrees) one-dimensional vector diagram presented in Figure 80.Although cycle analysis is continuing to investigate tradeoffs for a range of design "intturbine exit swirl, designing the rotor at the high swirl configuration allows added flexi-bility of tailoring exit swirl by trailing edge cutback as the analysis continues,
Preliminary stator vane profiles have been investigated for a range of vane numbersfrom 17 to 23, The final vane number selected will be based on rotor blade vibrationanalysis results, Preliminary vibration analysis indicates a stator vane number between19 and 21, A stator vane zµsection ring for a 19 vane configuration is presented in Figure81, The 2-dimensional vane surface velocity distribution for a parallel endwall ispresented"in Figure 82.
The objective of the vane profile design is to maximize stator inlet loading and mini-mize exit loading from the throat region to the trailing edge, This approach will mini-mize vane loading and, therefore, secondary flow losses when the contoured endwall isapplied. The effect of ,stator endwall contouring is shown in .Figure 83 for a 19-vaneprofile with reduced thickness, The symetrical endwall contour shown is similar to aprevious design when a 50 percent reduction in stator loss was predicted, For tilt, ACT,non-symetrical endwall configurations will also be examined to evaluate the effect onturbine inlet bend contour, The overall objective will be to optiunize the turbine inlet-stator systen'a,
In discussions with ceramic manufacturers, questions arose concerning stator trailingedge thicknesses and vane fillet radii, To investigate the effects of manufacturability onturbine performance, a trade-off study was conducted with varying fillet radii and trail-ing edge thickness. Figure 84 illust • tes results of this analysis as a function of vanenumber and trailing edge nori al thickness (based on axial turbine stator test results),As indicated, two configurations have been selected for further study, Configuration No,I is the AGT engine goal since no additional performance decrement is .indicated, Con-figuration No, 2 is more favorable from .a manufacturing standpoint and will, therefore,be designed and rig tested to verify the performance penalty associated with the in-creased trailing edge blockage.
Two other aspects of the profile design (maximum thickness and a straight line pres-sure surface) are also being investigated aerodynamically. From a thermal stress stand-point, a constant thickness profile would be ideal, Unfortunately, since the vane turningis over 70 degrees and the trailing edge thickness may be on the order of 0,508 mm(0,020 inch) and 0.762 mm (0,030 inch), maintaining a constant vane thickness is notaerodynamically feasible, However, a more uniform profile thickness can be achieved byreducing the vane maximum thickness as shown in Figure 85, It is also desirable tomaintain a straight line pressure surface to facilitate ceramic tool removal, Therefore,once acceptable stator vane velocity distributions are obtained, the change in loadingwith a straight line pressure surface will be examined.
112
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113
Q 21 VANES
0 1.9 VANES
TN • VANE TRAILINGEDGE THICKNESS
I/BT s BASELINE
EFFICIENCY
0.16 TN = 0.762 mm
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0.50 0.75 1.00 1.25 1.50 1..75 nun
L_ 1 i 1 1 f
0.02 0.03 0.04 0.05 0.06 0.07 IN
STATOR FILLET RADIUS, R
Figure 84. Effect of Stator Trailin g Edge Thickness and FilletRadius on Stator Exit Blockage and Turbine Performance.
114
MAXIMUM THICKNESS a 5.08 mm (0.200 IN )
TN = 0.508 mm (0.020 IN )
O—STRAIGHT LINE PRESSURE SURFACE
8 8.9 tan(3. 5 IN )
R 70. 87 mm(2. 719 IN )
THICKNESS - 3.81 mm (0.150 IN )
i .
R = 88. 9 mm(3. 5 IN )
i -R 70. 87 mm
(2.79 IN )
TN = 0.508 mm (0.020 IN )
STRAIGHT LINEPRESSURE SURFACE
Figure 85. Turbine Stator Profile Designs (19 Vanes).
115
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2.1,11.2,3 Stator Mechanical Design
A 3-dimensional finite element analysis of the reaction bonded silicon nitride (Si3N4)radial flowpath stator was done for the steady-state condition at 1644K (2500'F) TIT,A temperature plat for the 3-dimensional model is shown in Figure 86. The stator wasanalyzed for a segmented configuration and stress results are shown in figure 87. Workis continuing on an integral configuration,
Maximum stator temperatures are largely governed by combustor discharge tempera.tures and are found to approaQh these valuer on the airfoil, Shroud temperatures con-tacting the combustor baffle are warmer than those contacting c;,:;lcr turbine shroudcomponents; The thermal model assumes that a gap was pr; sent over a substantial Ix>r-Lion at the stator shroud/turbine shroud interface fe ► minimize differen,-es in statorshroud temperatures,
2,1.11.3 Turbine Rotor Design
2.1.11.3.1 Ceramic Rotor Aerodynamic Design
As stated previously, three levels of rotor exit swirl have been evaluated, In addition tothe level of swirl at the rotor exit, two different blade number designs have been evalu-ated; 12 blades and 14 blades, Figure 88 illustrates the effects of blade number versusrotor hub blockage and efficiency. In addition, rotor hub principal stress is depicted a afunction of blade number, Since little efficiency decrement is associated with either 12 or14 blades, blade selection will be dependent on rotor stress and blade vibration analysis,
Turbine meridional flow path is presented in Figure: 89, As indicated, the interim me-tallie rotor configuration is achieved by reducing the rotor tip speed from 701 m/s (2300fps) to 640 m/s (2,100 fps) by decreasing tip diameter, The exdu,cer will also be cutbackfor the metallic design, if the -37.0 degrees exit swirl vector diagram is selected, to pro-vide a further reduction in rotor inlet relative temperature, Rotor blade velocity distribu-tions for the hub, mean and shroud streamlines are presented in figures 90, 91, and 92for the ceramics rotor based on the following ;.onditions;
e Maximum powere 14 bladese 39,37 min (1,55 in,) rotor axial length• -37,0 degrees exit swirl vector diagram• 100,000 rpme 701 m/s (2300 fps) tip speed
If 12 blades are selected, rotor inducer loadings would be more severe,
116
Figure 86. AGT Stator Sectional Temperature Display.
117
CERAMIC RADIAL FLOi
63.4 MPa (9200 PSMAXIMUM STRESS
ORIGINAL PAGE ISOF POOR QUALITY
S MPa00 PSI)
^. 4W aPSI)
rv. a eLca tjdd5 PST)
Figure 87. Segmented Stator Non-Integral Band Stres Analysis.
118
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234 34
^a 22 1 - 32 I(r0 6.. 9 MP ak
H(1. p KSI)
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193 J 28
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BLOCKAGE LEVEL INSPECXFXC SPEED.CORRELATION
COMBINED EFFECTS OFBLADE NO. AND ROTOREXIT BLOCKAGE
10 12 14 16 18 20
N B — NUMBER OF ROTOR BLADES
Figure 88.. Effects of Blade Number and Exit Blockage onTurbine Performance an d Rotor Disk Stress,
. 11^
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5 R t.,$8.9 mm (3.500 IN )90
3.0 80
R=70.866 mm (2.790 IN )— h=6.54Q5 nun (0.2575 IN 1
70 ^Ot = 701 m/s (2300 FT/SEC)2.5 CERAMIC ROTOR
U = 640 m/s (2100 FT/SEC)6 0 G6.9442 mm MTALLIC ROTOR
METALLIC TRAILING(2.6356 IN EDGE
2.0a 50
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R-21. 3538 nun
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0 10 20 3.0 40 50 60 70 nun
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Figure 89. Turbine Merldlonal Flow Path.
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2.1.11.3.2 Ceramic Rotor Mechanical Design
The 2-dimensional preliminary turbine rotor stress analyses have been completed.These: studies were iterative in nature and were used to evaluate centrifugal stress effectsof various aerodynamic and mechanical design parameters. Table 8 summarizes ceramicturbine rotor properties and operating conditions utilizing both sintered silicon nitrideand sintered silicon carbide material, Figure 93 illustrates stress isopleths for these twocandidate rotor materials, For the 2-dimensional analysis, blade stresses were purposelykept low to compensate for geometric stress concentrations which occur in the transitionregion between blade and disk.
TABLE S. CERAMIC TURBINE WHEEL PROPERTIES
SINTERED SINTEREDSILICON SILICONNITRIDE CARBIDE_
Number of Blades 12 12Density, g / cm 3 (lb / in3) 3.27 3,16
(0,118) (0.114)Modulus, mPa (Ib / in 2) 310 x 407 x
103 103(45 x (59 x106) 106)
Poisson's Ratio 0,22 0,14Weight, kg (lb) 0,62 0.60
(1.36) (1.32)Ip Nm-s2 (lb-in-sec 2) 0.00037 0.00037
(0,0033) (0,0033)
Operating Conditions at Maximum Power (100,000 RPM)Inlet Temperature, K (° F) 1644 1644
(2500) (2500)Tip Speed m / s (ft / sec) 701 701
(2300) (2300)Average Tangential 151 145Stress, mPa (ksi) (21.9) (21.1)
After Completion of the 2-dimensional preliminary stress studies, three dimensionalstress studies were initiated to better understand stress in the blade/disk transition re-gion, Figures 94 and 95 illustrate the 3-dimensional finite element model. By utilizing
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symmetry, it is necessary to model only one blade and a corresponding segment of thedisk. Figure 96 illustrates good agreement between computed stresses for the 2- and 3-dimensional finite element models. The 3 -dimensional stresses in this figure are comput-ed at the center of the blade,. Consequently, these do not include any stress concentra-tions in the blade/disk transition region. Zoom modeling techniques are being utilized toevaluate peak stresses in this blade root region,
Blade natural frequencies have been calculated for the ceramic turbine rotor, TheCampbell diagram illustrated in Figure 97 i:, based on centrifugal effects using siliconnitride material, The Campbell diagram indicates no expected interference within theoperating range with any low .frequency excitations or with the stator passing frequencyif either 19 or 21 stator vanes are used. Figurc y 98 and 99 show mode shapez for the firstfour blade natural frequencies.
2.1.11.3.3 Metallic Rotor Design
Metallic turbine rotor definition was accomplished by first matching the ceramic nom-inal swirl (^27 degrees rotor exit swirl) at 1422K (2100°F) over the entire engine dutycycle, This established the turbine off-design speed and pressure ratio characteristics fora design tip speed of 701 in/sec (2300 ft/sec) utilizing ceramic, Since metallic rotor lifeis highly dependent on rotor relative temperature, rotor inlet relative temperature distri-bution, as a function of engine speed, was calculated from off-design turbine vector dia-grams as shown in Figure 100, Since metallic rotor turbine inlet: temperature (TIT) wasmaintained at 1422K (2100°F) from maximum power to idle, the turbine inlet relativetemperature increases as the rotor rotational speed is reduced below 80,000 rpm,Mechanical analyses indicated that under these conditions, a reduction in tip speed from701 m/s (2300 ft/sec) to 640 m/s (2100 ft/sec) is required to maintain an acceptablelife for the interim metallic rotor, This reduction in tip speed will be accomplished byreducing the rotor inlet, tip diameter, thereby maintaining maximum engine speed of100,000 rpm, Aerodynamically the effect of reducing rotor tip speed increases inducerincidence. Figure 101 shows a decrease in predicted overall turbine efficiency of approxi-mately l point due to tip speed limitations,
Table 9 presents mechanics! properties of the dual alloy metallic rotor, Figure 102illustrates the resulting metallic rotor stress field at maximum operating conditions,
2.1.11.3.4 Cold Turbine Test Rig
The cold turbine test rig, shown in Fig-ire 103, is an overhung design supported by twoangular contact ball bearings, Power section ducting and flow geometry have been repli-cated,
Rotor dynamics analysis has been accomplished, results of which are depicted inFigures 104 and 105. The cold turbine testing operating envelope is shown on Figure
126
Figure 94. AGT Ceramic Blade and Disk Segment.
4
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127
Figure 95. AGT Ceramic Blade and Disk Segment.
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Figure 97. AGT Ceramic Blade Campbell Diagram.
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.Figure 101 ACT, Metallic Turbine, Effects of Tip Speed on Efficiency [Tin - 1422K 17100"F1lat 10(),000 RPM,
1;i 4
104, Using bearing stiffnesses of 4378 N/m (25,000 lb/ gin), the first and second ur-damped critical speeds are calculated at 10,000 rpm and 18,000 rpm, respectively, Thesemodes, shown in Figure 105, are predominantly rigid body modes, The third critical is arotor bending mode and is calculated at 80,0,0 rpm, providing 56 percent margin aboutthe operating envelope. Since the C earings are hydraulically mounted, AiResearch Phoe-nix experience shows that characteristics of the mounts successfully damp the first vindsecond modes. Therefore, turbine operation is considered safe throughout the operatingrange,
TABLE 9, METAL TURBINE WHEEL PROPERTIES
Number of Blades
12Material , Mar-M-247 (DS) / AstroloyWeight, kg (ib)
1.45(3.20)
Ip (ib-in-sec2(0.0081)
Operating Conditions at Maximum Power (100,000 RPM)inlet Temperature K (°F) 1422
(2100°F)Tip Speed m/s (ft / sec) 640
(2100)Average Tangential (69.7)Stress (ksi)Burst Margin 1.34
Rotor operating clearance will be determined using clearance probes of suitable size atthe inducer, exducer, and backface planes. Variations in clearance are accommodatedthrough shim stacks and can be changed on the test stand.
Fiowpath instrumentation will be locatgd at the following stations;
• Stator inlet total pressure and temperature probes
• Stator static pressure sensor at the inlet, throat, exit, and selected pressure and suctionsurface locations
• Rotor inlet static pressure sensors
• Rotor shroud static pressure sensors
• Rotor exit total pressure and temperature probes
• Exhaust diffuser instrumentation
• Regenerator low pressure inlet instrumentation
Instrumentation will be sized consistent with small stage dimensions and be of suffi-cient number and spacing to provide realistic average values for all parameters mea-stgred,
135
(155.2 KSI)
(►
75E(1:
Figure 102, AGT Metal Wheal 2 .0 Centrifugal Equivalent Stresses — KSI, 12 Blades, 100,000RPM.
136
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PS STATIC PRESSUREPT TOTAL PRESSURE (–_ rTT TOTAL TEMPERATURE
! DIFFUSER EXIT PLANEPT,►S SURVEY'S
Ixis
COMBUSTOR EXIT PLANE -a^TURBINE EXIT PLANE
- PS,TT,PT SURVEY'S —CAPACITANCE -PROBES
TURBINE EXIT PLANEL. DIFFUSER INLET PLANEPT ,PS SURVEY'S
REGENERATORINLET PLANEHOT WIRE, PS
DISCHARGIE'^a V..REGENERATORPLANE( EXIT PLANE
TT PS
Figure 103,, Cold Turbine Test Rig Layout.F
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HEARING SUPPORTS STIFFNESS
Figure 104. Turbine Teat Rig Critical Speeds.t
138
Figure 105. Turbine Rig Rotor Dynamics.
139
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2.1.12 Ceramic Rotor Process Definition
2.1.12.1 Materials Selection
The ceramic radial rotor is the longest,, lcad, highest-risk ceramic component. Designstudies indicate that hot-pressed silicon nitride (Si3N4) and silicon carbide (Sic) haveadequate properties for the AGT turbine roor, although they are difficult and expensiveto fabricate in the required configuration, Emphasis must be on developing acceptableproperties in ceramic materials that can be fabricated by net shape processes. Reaction-bonded Si3N4 has net shape capability, but does not have adequate strength and is notprojected to have adequate strength by 1983, thereby ;eliminating it from further consid-eration. Sintered Si3N4 and Sic, reaction sintered Sic (RSSiC); and HIP Si3N4 havethe potential for meeting project requirements within the development timetable,
Considering current levels of development, projected properties, net shape capability,and cost, the following materials and potential sources have been identified;
Sintered SicSintered Si3N4Sintered Si3N4HIP Si3N4Reaction Sintered Sic
CarborundumFordA Research Casting Company (ACC)AiResearch/BattellePure Carbon/Piritish Nuclear Fuels,Ltd. (BNFL)
The following paragraphs describe candidate materials, fabrication approaches, andsources for ceramic rotor process definition in more detail,
2.1.122 Sintered SIC
Sintered Sic from Carborundum appears to be a near-term material for process eval-uation, Currently available Sic is variable with average flexural strength ranging from310-448 mPa (45-65 ksi) with Weibull m ranging from 5-10 depending upon the batch.This is not adequate for a reliable radial rotor, Average strength of about 621 mPa (90ksi) with a Weibull to of 1520 is required, However, Carborundum has just completed amajor new facility and can now conduct processing under dramatically improved condi-tions, Specimens fabricated under these type conditions have had an average strengthapproaching 621 mPa (90 ksi) with an in of 12, It is expected that, with continueddevelopment, comparable or better strengths will be routine,
Much development will be required to achieve high strength in an actual rotor. Carbo-rundum has attempted to fabricate an integral radial rotor by injection-molding underthe DOE/NASA/Chrysler program. To date they have not been successful due to theinability to remove the injection-molding plasticiser from the thick hub section, Current-ly, Carborundum is trying to solve this problem by evaluating alternate plasticisers and
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removal procedures, Also being explored is the approach (of injection molding a bladering and bonding it to a hub prepared by a process not requiring a plasticiser, Excellentbonding has been achieved on a specimen and it appears this approach is feasible for aradial rotor,
2.1.12.3 Slnterod S13N4
A survey of technical literature for the past few years would suggest that, sinteredSi3N4 component technology has developed more slowly than SiC and is consideredfurther from producing a reliable radial rotor, However, unpublished radial rotorfabrication studies at AiResearch Casting Company (ACC) and Ford Motor Companyhave demonstrated that sintered Si3N4 is just as viable as sintered SiC,
During the past two years, ACC has fabricated over 100 radial rotors of R,BSN usingexisting tooling of a turbocharger design, About 50 of these were slip cast to let shape,requiring only minor shaft and balance machining, Recently, the same tooling and Cast-ing procedures have been successfully used to fabricate sintered Si3N4 rotors using GTESylvania SN 502 powder,
ACC work to date has demonstrated manufacturing feasibility and has defined someof the tasks that would be required to develop a reliable rotor, Major tasks will be tooptimize the material composition to achieve required properties (strength, stress rupturelife, and oxidaticil-resistance in particular) and to eliminate procewsing flaws resultingfrom casting and sintering procedures,
Ford has also matte significant progress in sintered Si3N4 technology that has notbeen published. Using their own high purity Si3N4 powder with suitable sintering aidsadded, Ford has injection-molded the 820 engine integral stator configuration and suc-cessfully sintered it to 99 percent of theoretical density, More recently, Ford hasfabricated a radial rotor by slip casting and sinteree, it to 95 percent of theoretical densi-ty, Best strength values achieved to date for test specimens are a characteristic strengthof 614 mPa (89 ksi) with an m value of 8 at room temperature, and 434 mPa (63 ksi)with an in of I i at 1473K (2192'F), Although considerable development work remains,both strength and manufacturing feasibility have been demonstrated,
2.1.12.4 HIP S13N4
AiRextiarch has conducted an in-house program on Hot. Isostatic Pressing (HIP) ofSi3N4 for the past 1! years to evaluate HIP sources, explore Si3N4-Y203 compositions,and develop net-shape preform capability, From this study it, appears that HIP is a via-ble approach for a high strength, low cost radial rotor,
Hot isostatic pressing of Si3N4 is in the early stages of development in the UnitedStates, Battelle has demonstrated full densification of Si3N4-MgO and Si3N4-Y203
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compositions using Tantalum encapsulation. Glass encapsulation is required for low costand complex shapes, Battelle is currently developing glass encapsulation technologyunder a DARPA/AFML program,
The most extensive fi ll) of Si3N4 work has been conducted in Sweden at ASEAinvolving glass encapsulation to successfully fabricate individual airfoils and an integralaxial rotor. Currently ASI?A technology is not available,
Discussions with Battelle Columbus Laboratories have resulted in the preliminary def-inition of a HIP program directed toward radial rotor fabrication, Three approacheshave been identified as follows:
• Densification of powder preforms containing appropriate sintering aids• Densification of reaction bonded Si3N4 shapes containing a suitabet, sintering aid
• final densification of sintered Si3N4 or Sit;
The initial task of this program would be to optimize compositions in the Si3N4-Y?03, ` ON4-Y203-Si02 , and Si3N4-AL203 systems in terms of' densification andproperties. As a first step simple test bar shapes would be HlPped, At a later date, theshape would change to a simulated rotor having a mass comparable to the fully bladedrotor and finally, the actual rotor shape would be pressed. Parallel to this approach,reaction bonded Si3N4 containing Y203 would be HIPped, as would sintered rotorsfrom ACC, fiord, and C'arborundum, The best approach would be determined by spintesting.
Although initial HIPping of simple test. bar shapes could utilize Tantalum for encap-sulation, a fully bladed rotor will require that a suitable glass encapsulation technique beavailable. For this reason achievement of a suitable encapsulation glass has been identi.fied as a critical development area.
2.1.12.5 Reaction Sintered SC
Reaction sintered SiC (RSSiC) is also a potential rotor material, although not asmuch preliminary work has been conducted as on sintered SiC; and Si3N4. Some poten-tial advantages of R.SSiC over sintered mat.c:nals are;
• No shrinkage during densification
• Years of prototype and production experience
• Higher thermal shock resistance and fracture toughness
• Greater ease in incorporating fiber reinforcement
• Greater ease in fabricating a solid-hub design
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REPEL SiC from Pure Carbon/British Nuclear Fuels, Ltd (BNFL) is currently thebest RSSiC commercially available, Strength of pressed or injection molded REPEL istypically ovcr 516,7 mPa (75 ksi), Development has recently begun at BNFL on a slipcast version of REPEL, Average strength has been over 551,1 mPa (80 ksi) with individ-ual values over 689 mPa (100 ksi), Average strength over 620 mPa (90 ksi) should beachievable with further development,
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