Jet Engines Engine - Free Download PDF
Skip to content

Archives

Format Factory 4.6.0.2 License Key - Crack Key For U

Format Factory 4.6.0.2 License Key  - Crack Key For U

Couplings can also be selected in the product configurator of the X.CAT NG PC Full parallel key standard The parallel key is inserted in the shaft. Downloading Process: · Installation Process: · This software is Available With Crack full Version On? · if You Get/Went Free Trial Version OR Buy. Thus, high energy and capital costs, extensive factory manpower and floorspace, of the binder is below the activation temperature of the catalyst.

youtube video

Como achar o serial de qualquer programa

Format Factory 4.6.0.2 License Key - Crack Key For U -

818.Criminal Poisoning

Criminal Poisoning F O R SCIENCE- E N A N D S I C -MEDICINE Steven B. Karch, MD, SERIES EDITOR CRIMINAL POISONING: INVESTIGATIONAL GUIDE FOR LAW ENFORCEMENT, TOXICOLOGISTS, FORENSIC SCIENTISTS, AND ATTORNEYS, SECOND EDITION, by John H. Trestrail, III, 2007 MARIJUANA AND THE CANNABINOIDS, edited by Mahmoud A. ElSohly, 2007 FORENSIC PATHOLOGY OF TRAUMA: COMMON PROBLEMS FOR THE PATHOLOGIST, edited by Michael J. Shkrum and David A. Ramsay, 2007 THE FORENSIC LABORATORY HANDBOOK: PROCEDURES AND PRACTICE, edited by Ashraf Mozayani and Carla Noziglia, 2006 SUDDEN DEATHS IN CUSTODY, edited by Darrell L. Ross and Ted Chan, 2006 DRUGS OF ABUSE: BODY FLUID TESTING, edited by Raphael C. Wong and Harley Y. Tse, 2005 A PHYSICIAN’S GUIDE TO CLINICAL FORENSIC MEDICINE: SECOND EDITION, edited by Margaret M. Stark, 2005 FORENSIC MEDICINE OF THE LOWER EXTREMITY: HUMAN IDENTIFICATION AND TRAUMA ANALYSIS OF THE THIGH, LEG, AND FOOT, by Jeremy Rich, Dorothy E. Dean, and Robert H. Powers, 2005 FORENSIC AND CLINICAL APPLICATIONS OF SOLID PHASE EXTRACTION, by Michael J. Telepchak, Thomas F. August, and Glynn Chaney, 2004 HANDBOOK OF DRUG INTERACTIONS: A CLINICAL AND FORENSIC GUIDE, edited by Ashraf Mozayani and Lionel P. Raymon, 2004 DIETARY SUPPLEMENTS: TOXICOLOGY AND CLINICAL PHARMACOLOGY, edited by Melanie Johns Cupp and Timothy S. Tracy, 2003 BUPRENOPHINE THERAPY OF OPIATE ADDICTION, edited by Pascal Kintz and Pierre Marquet, 2002 BENZODIAZEPINES AND GHB: DETECTION Salvatore J. Salamone, 2002 AND PHARMACOLOGY, edited by ON-SITE DRUG TESTING, edited by Amanda J. Jenkins and Bruce A. Goldberger, 2001 BRAIN IMAGING IN SUBSTANCE ABUSE: RESEARCH, CLINICAL, edited by Marc J. Kaufman, 2001 AND FORENSIC APPLICATIONS, CRIMINAL POISONING INVESTIGATIONAL GUIDE FOR LAW ENFORCEMENT, TOXICOLOGISTS, FORENSIC SCIENTISTS, AND ATTORNEYS Second Edition John Harris Trestrail, III, RPh, FAACT, DABAT Center for the Study of Criminal Poisoning, Grand Rapids, MI © 2007 Humana Press Inc. 999 Riverview Drive, Suite 208 Totowa, New Jersey 07512 www.humanapress.com All rights reserved. No part of this book may be reproduced, stored in a retrieval system, or transmitted in any form or by any means, electronic, mechanical, photocopying, microfilming, recording, or otherwise without written permission from the Publisher. The content and opinions expressed in this book are the sole work of the authors and editors, who have warranted due diligence in the creation and issuance of their work. The publisher, editors, and authors are not responsible for errors or omissions or for any consequences arising from the information or opinions presented in this book and make no warranty, express or implied, with respect to its contents. This publication is printed on acid-free paper. ∞ ANSI Z39.48-1984 (American Standards Institute) Permanence of Paper for Printed Library Materials. Production Editors: Tara L. Bugg Cover design by Nancy K. Fallatt For additional copies, pricing for bulk purchases, and/or information about other Humana titles, contact Humana at the above address or at any of the following numbers: Tel: 973-256-1699; Fax: 973-256-8341; E-mail: [email protected]
Источник: https://www.docme.su/doc/1346503/818.criminal-poisoning
LUBE FILTER SCAVENGE

INLET SCREEN (TYPICAL)

AFT " C " SUMP

DISCHARGE \ SUPPLY \ D,SCHARGE

SCAVENGE INLETS M A G N E T I C CHIP DETECTOR (OPTIONAL)

FILTER SERVICE SHUT-OFF VALVE

MICRON 74 LUBE FILTER SUm.Y INLET

- f l r«j * ^

FILTER BY-PASS VALVE (40 PSIO)

STATIC LEAK Ctfl-: •$} CHECK VALVE j * PSID

SPLINE LUBE SUPPLY

Figure 8.17 Lube and Scavenge Pump

Lube Scavenge System The oil scavenge system consists of the gravity drain regions in the sumps including piping to the pump, scavenge pump, filter, and heat exchanger. The purpose of this system is to prevent oil storage in the sumps and transport the sump generated heat to an outside cooling source. Lube Scavenge Pump Scavenge pump elements can be of the same types as supply elements but are sized to have three times the capacity of the oil delivered to the sump being scavenged. This excess capacity assures dry sump operation so that bearings do not run submerged in oil which could cause excess heat generation. This excess capacity is the primary reason a deaerator is required in the lube tank. Scavenge elements should be primed by gravity. The elements are usually required by specification to be self priming to inlet pressures as low as 1.5 psia. The discharge side of the elements must be submerged in oil. Scavenge element seizure should cause failure of the supply pump as well to prevent the possibility of sump flooding. Fuel Oil Cooler The oil cooler regulates the temperature of engine oil by transfer of heat from the oil to the engine

SECONDARY SYSTEMS

fuel. The properly designed cooler thermodynamically responds to changes in heat load (oil flow and temperature) and heat sink (fuel flow and temperature) so as to maintain required engine oil temperatures. Heat transfer performance requirements for the oil cooler are established by systems analysis of the engine. Overall considerations of engine weight, available envelope space, cost, and reliability are also considered in establishing fuel oil cooler performance requirements. Limits of fuel and oil pressure drop are imposed on the cooler as well. Other important considerations are failure effects such as fuel leaking into oil or to external regions, maintainability (cleanability), and repair. Chip Detectors Chip detectors are magnetic devices installed in main or individual scavenge lines or in the bottom of oil tanks or gearboxes to collect wear particles. These devices all collect particles by means of a magnet and are useful for collecting chips greater than 50 micrometers. Also available are electric chip detectors with connections for remote (cockpit) indication. Several varieties of magnetic chip detectors are shown on Figure 8.19.

8-19

( l o l l VANE PUUP

K^f

CAPTURE EFFICIENCY 5 TO 14%

^7 / (A)

GEAR PUUP

FLOW IN

15 TO 25% FLOW OUT

(B)

INTERNAL GEAR PUUP 10D% (PARTICLES LARGER THAN SCREEN MESH)

(C) Figure 8.18- Various Types of Pumping Elements

8-20

Figure 8.19 Chip Detectors

SECONDARY SYSTEMS

Sump Vent System Sumps can be either vented or nonvented designs. A schematic of a vent system is shown on Figure 8.20. The aft most sump on Figure 8.2 is an example of a non vented sump. The sump vent system provides the following functions: to maintain sump pressure below pressurization air levels, to maintain scavenge pump inlet pressure above the minimum acceptable performance level, to provide acceptable oil consumption rates, and to provide transient capability.

Non-vented sumps have quite different design requirements. These sumps must necessarily be small in volume and located in low pressure areas, usually in the forward or aft end of the engine. These sumps rely on the excess capacity of the scavenge pump to prevent oil leakage from the sump. Ambient temperature is limited to 600 °F. Vent flow is directed overboard usually through an airoil separator which removes oil particles from the air to minimize oil usage.

Air contained in the vent system is supplied by leakage across sump oil pressurization seals. If the vent circuit exiting the sump is oversized, the sump pressure will be low. This causes increased seal flow which in turn increases oil consumption by carrying more oil particles overboard. The low sump pressure also is detrimental to scavenge pump performance. If the vent circuit is too small, the sump pressure will be too high, forcing the oil out through the oil seals. The secondary systems designer must size these vent areas so that proper balance is maintained during both steady state and transient operation.

Air-Oil Separators Lack of adequate air-oil separator capacity has been the most common cause of high oil consumption on jet engines. Static separators usually mounted in lube tanks, have low air capacity, typically 30 standard cubic feet per minute (SCFM). Dynamic separators, usually built into a gearshaft or main rotor shaft, can have capacities up to 250 SCFM. Dynamic separators are used in most recent engines. A schematic of a gearbox mounted and intershaft mounted separator are shown in Figure 8.21. Both are examples of dynamic separators. AIR/OIL SEPARATOR

—OVERBOARD

OIL RETURN

Figure 8.20 Vent System

SECONDARY SYSTEMS

8-21

INTERSHAFT AIR/OIL SEPARATOR

AIR/OIL SEPARATOR OIL OUT

*- AIR OUT

AIR/OIL IN

6000 RPM Figure 8.21 Air-Oil Separators *-?->

SECONDARY SYSTEMS

Oil Consumption For engine designs with labyrinth seals, oil consumption is largely a function of vent airflow and temperature. Over the total life of the engine the seals tend to wear so vent flowrates gradually increase. For this reason, oil consumption quoted in engine specifications is usually a maximum guarantee level. Typical consumption numbers for a variety of engine designs range from .05 to .30 gallons per hour. Oil Filtration Serviceable contamination filters should be provided in judicious locations throughout the lube system. Coarse filters for items such as weld splatter or machining chips should be backed-up by fine filters for materials such as sand or engine wear particles. The fine filter is the "working" filter which establishes the system cleanliness level. This method should provide long reliable lube system life. Prior GE experience has been with fine filtration in the supply system alone, scavenge system alone, and in both supply and scavenge systems. The advantages of fine filter in scavenge system alone are wear particles are removed closest to the point where generated, filter debris shows up sooner for trouble indication, and the filter provides engine contamination protection for cooler, tank, anti-staticleak valve and supply pump. Advantages of fine filter in supply system alone are that it provides filtration immediately before distribution to oil-supplied components so that both engine generated debris or possible contamination from tank servicing is eliminated. If circumstances dictate the use of only one filter, it must be the supply filter that is used on the basis of maximum component protection. However, design practice is to provide both a coarse supply and a fine scavenge filter. Lube Heat Rejection One of the major functions of the secondary systems engineer is to calculate the heat load of the oil system. Estimates of heat transferred to the oil by bearings, seals, sump walls, pressurization air, pumps and gears must be made for extreme ambient air and extreme fuel temperatures expected during in service operation. Typical results of these estimates are shown on Figure 8.22. This data is then used to estimate the oil heat load effects on the fuel system to prevent associated problems such as fuel nozzle coking. Once the analysis has been completed, a computer model of the system is constructed in a computer language compatible with airframe companies systems. Airframe designers have pumps and cooling equipment that interact with the fuel and need our input to optimize their systems.

SECONDARY SYSTEMS

Fire Safety Analysis Secondary systems engineers have the responsibility to analyze the potential for combustion in the sumps and surrounding areas. The analyses must include normal operation and any malfunction of systems or hardware. Assuming that a problem, or potential problem, has been isolated to a particular flow circuit and fuel source, the following process is applied: obtain information on suspected operating condition and location, obtain drawings and estimates of system flow, pressure and temperature, split the flow into a series of chambers, and analyze each camber for conditions of flammability, ignition and stability. Flame can be either pre-mixed, which is a homogeneous fuel vapor/air mixture, or of the diffusion type which is a localized liquid fuel puddle source. In each case, all conditions of flammability, ignition and stability must be met. This method provides either a zero or 100% probability of fire. These methods are explained in detail in reference material.

SUMP/SUPPORT HEAT TRANSFER ANALYSIS Accurate prediction of metal and air temperatures for bearings, sumps and surrounding cavities is critical to the design of these components. Setting bearing operating clearance, determining air/oil seal clearances, estimating axial travel of rotor relative to stator, and stress analysis are four major uses for the data. Detailed heat transfer in the bearing region is extremely difficult due to the complexity of the oil flow regime in and around the bearings. A cross section of one of the more complicated bearing and support models is shown in Figure 8.23. This 1100 node model was used to predict steady state bearing, seal, and support temperatures plus transient temperature changes in seal and support structure. Component rig testing was used to determine the bearing heat generation characteristics for input into the model. .. < Once a complete nodal model of the system was available, secondary systems air and oil models and engine test data were used to calculate flows in sumps and surrounding areas. These flows were "then used to calculate heat transfer coefficients. GE's heat transfer program, THTD, was run to calculate local metal temperatures for comparison to measured engine data. The heat transfer coefficients were adjusted as required until a temperature match was obtained. Boundary temperatures and pressures used in this analysis were normalized by major

8-23

375

ALT - 35K RP - .20

350 325

u. o

WN

^

300

UJ

, H ^

250

UJ

^

•B1 1

- 130° F FUEL - 85° F FUEL - 60° F FUEL - 30° F FUEL 0° F FUEL

,r-A

\

225

=> < ce

*o-

I

275

D O A + x

^

200

Y~-

175

UJ D_ X UJ

150 125 60 65 70 75 SO fl5 90 95 100 105 110

'PERCENT PHYSICAL CORE SPEED

375

GROUND RP - 0.0

350 325

o. o UJ DC

300

r

275 250

-*

Q:

200

UJ CL

21 UJ

a

0 A +

B at

X

a

V

130° 85° 60° 30° 0°

F F F F F

FUEL FUEL FUEL FUEL FUEL

/ > /, ^

225

<

r^

^

a

V.

t

j

s

175

, ^^

150

/

>

125 20 30 40 50 60 70 80 90 100 110 120

PERCENT PHYSICAL CORE SPEED

Figure 8.22 Calculated Lube Scavenge Temperature For Various Fuel Temperatures

8-24

SECONDARY SYSTEMS

FLUID ELEMENTS

Figure 8.23 Forward Sump Heat Transfer Model

SECONDARY SYSTEMS

8-25

engine parameters such as engine inlei or compressor discharge temperature so predictions for cycle points could be made. The nodal model for this effort is shown in Figure 8.23 for the b-sump. A similar model exists for the c-sump. Using these models, metal temperatures for particular sump components were predicted at cycle points requested by mechanical designers and the results forwarded to them for use in their design effort. Another very important heat transfer function, particularly for new engine designs, is to estimate the maximum oil wetted metal temperature expected for a sump configuration. Lubricating oils currently in use tend to generate vamish and start to form coke in the 440 °F to 450 "F temperature range. Severe coking can cause plugged lube supply and scavenge circuits, cause carbon seals to hang up or just be a general maintenance or cosmetic nuisance. ,. ,. Design practice is to limit oil wetted metal temperatures to 400 °F. This is easier to accomplish during engine operation than during post shutdown periods when all the heat stored in the engine tends to radiate to the sumps with no flow mechanism to cool the sumps. This phenomenon is known as soakback. If not designed for it can drive sump wall or tube temperatures above coking levels.

AXIAL BEARING THRUST CONTROL Balancing the.axial thrust loads which develop in the flowpath and internal cavities of jet engines is critical to obtaining acceptable thrust bearing lives. Since the secondary systems air model includes all major cavities, a simple summation of pressure times projected area for all pertinent cavities will give the resultant load on all rotating hardware. Factoring in the compressor and turbine aerodynamic blade loads yields the axial forces on the engine thrust bearings.

8-26

HP Rotor Thrust Table 8.1 shows a schematic and tabulation of pressures, areas, and forces involved in obtaining me resultant loads on the high pressure rotor bearings of the CF6-80C engine at takeoff. Note that the resultant load (-4561 lb.) is small relative to the major cavity loads. This is typical of HP rotor bearing axial loads. Also, note that maximum cavity loads are substantially higher than either compressor or turbine total airfoil aerodynamic loads. Compare forces 104 and 114 with force 113 on Table 8.1. The accuracy of predicting these cavity pressures is a critical factor in predicting bearing loads. Once the bearing load is predicted and determined to be too high for acceptable bearing life it can usually be adjusted to required levels by moving a critical seal to a larger or smaller diameter thus changing its projected area.. For instance, in Table 8.1, changing the diameter of the seal which affects forces 105 and 106 would be used to balance the load on the HP thrust bearing. In drastic cases if more adjustment is necessary, several seals or even turbine airfoil changes may be required to obtain desired bearing loads. Figure 8.24 shows a comparison of four different engines axial HP load and how they change with engine speed. LP Rotor Thrust Prediction of the low pressure rotor thrust bearing load is generally much easier mainly due to the lower pressure levels involved. The principles remain the same. Figure 8.25 shows the low pressure rotor thrust for the same CF6-80C engine. Note that the predicted load is very close to the measured data presented in the figure.

INTERFACES Performing the above design functions necessarily involves interactions with many other specific groups internal and external to the company. Figure 8.26 is a schematic showing typical types of functions or subjects involved on the radial lines and the groups interfaced with during those functions.

SECONDARY SYSTEMS

GO

m o O

> •<

w H

m

107 W

uttt&tuj.

NO 101 102 103 104 105 106 107 108 109 110

a

119

JSL

Rt 3 . 190 3.845 4.840 1 7.640 6.375 4.650 4.000 4.000 4.325

CHAM AREA PRESS FORCE RO 14.48 19.69 -285 3 . 8 4 5 30 27.16 33.07 -898 4.840 33 5 7 7 . 8 9 24.71 -1925 6.944 — — +5783 1 7 253.92 364.65 +9332 i1.825 5 5 . 7 0 102. 8£ +5727 7 . 6 5 0 36 5 9 . 7 5 2 2 . 7 5 +1359 59 6.375 1 7 . 6 6 2 6 . 6 5 +471 41 4.650 1,90 2 0 . 9 3 +40 58 4.075 2 1 . 3 5 2 1 . 3 2 -451 5 . 0 5 0 62

NO 111 112 113 114 115 116 J17 118 119

Rl

CHAM RO 39 5.050 6.200 40 6.200 7.470 10 7 . 4 7 0 4.04C — 1 1 2.89C 1 1 . 825 11 12 12.89CI4.00C 6 . 8 3 0 13.660 51 4 . 4 2 5 3 . 5 4 3 54 50 2.925 6.830

Table 8.1 Variables Affecting Loads On High Pressure Rotor Bearings

AREA 40.64 54.34 •48.39 85.04 93.77 64.23 429. 66 85.04

PRESS FORCE 105.14 - 4 2 7 3 316.95 -1130£ 253.99 -118879 — -31856 186.081-1744* 128.76 - 8 2 8 2 7 6 . 2 2 +41339 8 3 . 7 0 +7118 2 5 . 0 9 + 869

O^

1

-1

-2

-3

CORE SPEED RPM (THOUSANDS)



CF6-6

•f

CF6-50

O

CF6-80

A

CF6-80C

Figure 8.24 CF6 HP Axial Thrust Comparison

8-28

SECONDARY SYSTEMS

-

2

o CO CD

0

-14

8 0

8

16

24

32

Источник: https://kupdf.net/download/jet-engines-engine_5af44dd6e2b6f5ec7b54c8d1_pdf
0

1 .1

I I

A

i

-. -j 1 - - -_ 1X ™_ _Hi

1i

1 .2

.22

PT/PT

.24

,--iq-

1 1 i i

50,

3ol--

loo,

n OPEN CLEARANCE OUGHT CLEARANCE

.28

AB5. FLOW ANGLE

SINGLE-STAGE HPT EXHAUST PRESSURE AND FLOW ANGLE DISTRIBUTIONS FOR TWO LEVELS OF T I P CLEARANCE

Figure 5.21 Single-stage HPT Exhaust Pressure and Flow Angle Distributions for Two Levels of Tip Clearance

CC DR

X lC0

r> < 3

tUJ

o _l u.

_J

H 10 O < X X UJ

CC

z o oCC t-

# u z

01

o

tCO

o m o u u X

• Figure 5.36 Static CDP and FOS Seals 5-36

TURBINES

HP TURBINE BLADE

AFT AIR SEAL

COW-RE SSOR DISCHARGE PRESSURE SEAL ECDPJ

Figure 5.37 HP Turbine Motor

Compressor Discharge Seal (CDP) Disk This component is sized to provide a seal for compressor discharge air as discussed earlier. A disk is required to support the rotating seal head and to keep the concentrated stresses in the bolted flange to levels low enough to avoid low cycle fatigue. Another factor in the size and proportion of this disk is to keep critical rotor vibratory frequencies out of the range of operating speeds. Forward Shaft The purpose of this shaft is to provide a torque and axial load carrying structural member between the HP disk and HP compressor. Again this component is proportioned to avoid critical rotor vibratory modes and low cycle fatigue. Forward Outer Seal (FOS) Disk and Retainer The rotor disk is relatively complex as disks go. As with the CDP seal disk, it provides support for the second (outer) cavity sealing head, sometimes called the "four tooth seal". Cantilevered off of the seal head is a flexible arm supporting the forward blade retainer. This structure provides assurance that the HP blades will not move forward in the flowpath. The retainer portion of this

TURBINES

structure is rabbeted to me HP disk. Rabbeted construction has three basic purposes. It provides additional radial support as the FOS arm could not withstand bending stresses associated with centrifugal loads. Secondly, this construction assures that rotor parts do not shift under heavy centrifugal loads. Shining would destroy the impeccable balancing of these high speed components. Imbalance associated with shifting of rotor parts would cause heavy one per rev vibratory loads, generate excessive seal leakages, and could generate high cycle fatigue. Finally, rabbetted construction provides a seal for the blade cooling circuit. Blade cooling air would be dumped into the FOS cavity without this seal (see Figure 5.38). Besides being a performance loss, without the cooling air the HP blade would overheat and rupture. The web of the FOS disk has holes placed slightly below the inducer center line. These holes permit the inducer to pressurize the blade cooling circuit. Remember the inducer accelerates the blade cooling air to meet these rotating holes and thereby expanding (cooling) the air. Since the air is now moving at wheel speed it can get on board the rotor without heat being generated through windage. This is a double benefit: less drag on the rotor (more efficiency) and increased cooing capacity for the

5-37

Figure 5.38 Secondary Air Flow HP blade. Placing these holes below the center line of the inducer helps keep dust out of the blade cooling circuit. The air can turn the corner and the dust if it is heavy enough will be centrifuged back to the flowpath. Radially, inboard of the inducer holes, is a two tooth seal which prevents blade cooling air from leaking into the CDP cavity or hot CDP cavity air from leaking into the blade cooling circuit as a function of pressure gradients. The FOS disk bolts lo the HP disk and forward shaft. The complex shape of the FOS disk can result in unnecessary thermo-mechanical bending stresses. When these features are carefully positioned with respect to each other these stresses can be minimized. This is called "stacking the disk". A full understanding of the time varying thermal gradients and mechanical loading is the key to successfully "stacking the disk". HP Disk This component supports the blade dead load (over 2 million pounds), along with the radial rabbet loads from the forward and aft retainers. The HP blades are held by a two tanged fir tree arrangement shown in Figure 5.39. The large radial blade load and disk mechanical and thermal stress result in a very large bore

5-38

hoop stress. This stress is reduced by increasing the bore width (see Figure 5.37). Disk flange arms facilitate bolting other shafts and disks to the HP disk. Again this disk is proportioned to avoid critical vibratory modes in the engine operating speed range. Aft Retainer This component is rabbeted to the HP disk and is charged with providing an axial stop for the HP blade. It also provides sealing for the HP blade cooling circuit. When forward and aft retainers seal properly, the blade cooling circuit is pressurized by the inducer and the hollow HP blade is convectively cooled. Aft Shaft This rotating structural member provides support for the HP rotor as its journal houses a radial bearing. The shaft also meters the bore cooling circuit. Shaft holes permit bore cooling air to leave the HP rotor and pass on to the LP rotor, (see Figure 5.38). As with the CDP and FOS disks, this shaft has seal teeth which interact with the static and rotating low pressure (LP) turbine, thereby sealing HP and LP secondary flows from each other. Again, this component is proportion to avoid critical rotor vibratory modes in the operating speed range.

TURBINES

80 DOVETAILS

72 DOVETAILS

CFM56-5

CFM56-3

Figure 5.39 Comparison Between CFM56-3 and CFM56-5 Dovetails

HP Blade The HP turbine has eighty hollow cast turbine blades. The internal passages have been designed to provide for an efficiently cooled turbine airfoil. These circuits are shown in Figure 5.40. Convection is the prime heat transfer mechanism. After the cooling air speeds through the blade passages, it is returned to the flow path gas stream through small holes in the airfoil and tip cap.

The blade, like the turbine disk, must be stacked carefully to reduce bending stresses caused by gas path pressures and centrifugal loads. The aggressive thermal environment can even cause tough blade alloys to oxidize or corrode away so the blade receives a CODEP coating to protect parent blade metal.

THE MECHANICAL DESIGN PROCESS

Air exiting the airfoil forms a thin film of air around the blade to minimize disruption to gas streamlines and to shield the external blade surface from hot flowpath gases. Thin trailing edges and gas path stagnation at blade leading edges cause these locations to be the hottest. Airfoil holes are usually placed in these locations with design intent being to minimize airfoil thermal gradients.

The design of HP turbines is a challenging problem, that has very little margin for error. Complex thermomechanical stresses generated in engine operation can produce failure in meeting design goals and sometimes actual engine disfunction. The road to a successful design is attention to detail. Over looked items usually lead to serious problems.

There are three functions for the air exiting the blade tip. The first function is to cool the tip cap, the second is to purge internal blade circuitry and reduce internal pressure loss, and thirdly it purges the blade of dust. If dust accumulates in blade cooling passages or blades cooling holes the result will be a hot blade and rupture/fatigue.

While the dream of every turbine designer is a clean sheet of paper, the reality is a nightmare of deciding where to start. Then, there are seemingly endless iterations of refinement and discovering how to compromise the conflicting design goals and constraints of weight, performance, and durability.

TURBINES

5-39

TIPCAP PURGE HOLES

CAVITY I CROSSOVER HOLES [IMPINGEMENT COOLING)

LEAD EDGE COOLING H X E S

PITCH SECTION

TRAIL EDGE COOLING HOLES

3 CIRCUIT HPT BLADE COOLING Figure 5.40 Cooling Circuitry for CFM56-5 HPT Blade

Whiie flowpath hardware (nozzles and blades), as well as stator and rotor hardware each have their own particular design problems. They also have similarities. To illustrate the process let us focus on the rotor to get a sense of how a mechanical design evolves. An overview of design and development includes the following activities: Basic design concept Preliminary sizing Engineering drawings Working the details Component and engine testing, and Final certification or qualification analyses All designs rely on four critical items, teamwork, experience, demonstration, and verification. A turbine mechanical designer is a conductor of a large orchestra composed of specialist in areas like: Cycles and performance Aerodynamics Secondary flow

5^0

Heat transfer Materials application Drafting Testing Computer methods Stress and vibration analysis, and Manufacturing The designer must meld the design inputs and constraints from each of these areas into a harmonious result. Experience is found in a variety of places technical specialists, other mechanical designers, GE design practices, reviews held by the Chief Engineer's office, and individual designer's own personal reservoir of knowledge and insight. No matter how strong the designer's knowledge and insight is, it will be insufficient to producing a viable design. Experience in applying engineering principles will result in a sound design concept, only when all resources are utilized in a detailed fashion. A design concept can be refined and verified through repeated modeling and analysis. This is an integral portion of design verification. Application of design analysis

TURBINES

methods tempered by lessons learned is also a powerful cost effective method for accumulating and applying actual design experience. Final demonstration of all designs is achieved through component and engine testing. The engine always processes a level of wisdom that humbles the best thought out designs. The Basic Design Systems engineering has prime responsibility for the total engine system meeting its requirements. Their relationships with cycles, performance, aerodynamics, and mechanical design specialists gives broad form to the engine and its requirements. Key items here are: Number of turbine stages Flowpath dimensions Aerodynamic shapes for the nozzle and blade Numbers of nozzles and blades Engine speed and thermodynamic parameters, and Locations of frames and bearings Once the size and performance parameters of the HP turbine are established by systems engineering the mechanical design process can begin.

Preliminary Design The core engine rotor structure is usually supported at its extremes by bearings, one of which is an axial thrust bearing. (Figure 5.41). For the CFM56/F101 family of engines the forward bearing supporting the HPC rotor, is the axial thrust bearing. This bearing supports the net axial load on the core rotor. The sources of this load are HPC and HPT blade loading and internal pressures acting on the rotor structure. This bearing is also the reference point for axial growth of the rotor, an important factor in establishing rotor seal design. With the inputs provided by Systems Engineering, the mechanical designer begins discussions with Systems, Secondary Flow, Heat Transfer, Bearings Design and Performance specialists to define the secondary flow aspects of the turbine structure. Fundamental engine issues are being settled here. Major issues include: How much bleed air will be used to cool and purge each rotor-stator cavity and the HP turbine blade? Where will the bleed air be extracted? How will the bleed air be reintroduced to the engine flowpath?

STAGE 1-2 SPOOL

FWD. SHAFT

COMPRESSOR RADIAL BEARING

THRUST BEARING

Figure 5.41 CFM56-3 High Pressure Rotor Bearing Support

TURBINES

5-41

Given the above, what is the impact on engine performance?

is generating engineering drawings to support technical work.

How can engine performance be improved?

An early milestone is the material release drawing for forgings and castings. Lead time for these items can be from six months for forgings to one year for castings and must be allowed to achieve the overall engine schedule. Experience, intuition, and engineering estimates are used to provide detail component definition for fillets, hole size, and other features which can result in significant level of concentrated stress.

What will be the diameter of the compressor discharge and the pressure balance (forward outer) seals? Is the resultant axial load within the bounds adequate for bearing life? Does this secondary flow system present insurmountable mechanical problems? Resolving these fundamental issues requires creative thought, evaluation of many alternative concepts through an iterative process, and then arriving at a design based on mutual compromise. During this process, estimates of internal cavity flows, pressures and temperature are made by secondary flow specialists. The heat transfer specialists estimate metal temperature distributions for generic acceleration from steady state idle to steady state takeoff and a deceleration back to idle. With these estimates and engine speed (a thermodynamic parameter) the mechanical designer begins to size the various turbine rotor disks. Disks are usually required to support rotating seal heads and large dead loads like those created by the rotating turbine blades. The designer must assure that each disk has adequate overspeed margin to its burst speed, and then estimate radial and axial deflections for rotating seal heads as well as the HP turbine blade/ These displacements when compared to deflection estimates made by those designers responsible for the HP stator provide insight to potential seal clearances. These clearances usually require the secondary flow analysis to be repeated, and so on until the rotorstator interaction is understood and secondary flow and performance constraints are satisfactorily met. This process requires many months of work until the final design concept is worked out. Engineering Drawings During this early period the mechanical designer will work out associated details relative to sizing shafts and bolting issues. In sizing each HP turbine component the designer must also assure that adequate materials are selected for each component, operating bulk temperature and stress distributions are kept to levels consistent with creep, rupture, and low cycle fatigue life requirements, disk and rotor vibratory modes are kept outside the engine operation range, and drafting

5-42

After much design work and discussion, involving design and design analysis reviews with the Chief Engineer's office and other experienced mechanical designers, the final engineering drawings are released for each component. Again, timing is important due to manufacturing lead time. A sufficient period is required for tooling, gaging, and machining. Typical process requirements are from three to six months. Working the Details From the release of engineering drawings to the first chips made cutting the initial turbine components, a time period is provided for detailed design analysis of the HP turbine rotor. Early in the engine program. Systems Engineering has defined a design flight cycle or cycles for engine wide application (Figure 5.42 and Table 5.4). This operating cycle is a generic description of revenue service engine usage. It is used by design and component systems specialists to establish the transient and steady state turbine operating environment. Rotor heat transfer specialists construct a transient heat transfer model of the rotor (Figure 5.43). The resulting transient metal thermal distribution shown in Figure 5.44 is for the "end of takeoff mission point. This model is based on steady state parameters determined by secondary flow specialists for the design flight cycle (Figure 5.45). Mechanical Design specialists then generate a structural rotor model (Figure 5.46) composed of shells, rings, and two dimensional finite element members. Transient thermal distributions are then input,to the rotor model for approximately 100 time points of interest. This structural model provides basic understanding of bulk stress behavior of the rotor with time (Figure 5.47). To determine levels of concentrated stress around fillets, holes, slot bottoms, and complex non-symmetric rotor features more detailed models are required. Now, two and three dimensional finite element models are generated to fully describe component geometries and predict correct values of concentrated stresses. Shown in Figure 5.48 is the bolted joint where the forward shaft and

TURBINES

]2

J±4

THRUST REVERSE

CD

CO

(X UJ O Q_

TIME Figure 5.42 CFM56-5 Design Flight Cycle Profile and Points

ALTITUDE (FEET)

POWER

MACH NO.

L/M/N/O* DTAMB (°F)

INTERVALS OF TIME (MIN) 1.39

0

.35

-9/11/27/66

1500

.40

-9/11/27/66

10000 20000 30000

.61 .733 .850

10/15/20/40 10/15/20/40 10/15/20/40

20.0

CRUISE

30000

.850

10/15/20/40

19.24

DESCENT

30000 20000 10000

.859 .733 .610

10/15/20/40 10/15/20/40 10/15/20/40

14.0

5000

.3

10/15/20/40

5.0

FIDLE

0

.2

-9/11/27/66

0.11

THRUST REVERSE

0

.2

-9/11/27/66

0.26

GiDLE

0

0

-9/11/27/66

5.0

TAKE-OFF END TAKE-OFF CLIMB

APPROACH

- RELATED TO STD DAY

c o o >

a

n m o

25 o z

SUB-SYSTEMS

COMPONENTS ACCY DRIVE INTERNAL G/B ACCT DRV GRZ BEVEL GEAR ACCT DRV RAD DRV SHAFT ACCY G/B HORIZ SHAFT ACCY G/B HRZ SHAFT HSG A/C ACCESSORY GEARBOX AUGMENTOR MIXER AUGMENTOR DUCT AUGMENTOR FUEL MANIFOLD AUGMENTOR FUEL TUBES AUGMENTOR FUEL VALVE AUGMENTOR IGNITOR AUGMENTOR LINER AUGMENTOR SPRAYBARS COMBUSTOR (IGNITION) COMBUSTOR CASE COMPRESSOR BLADE/VANE COMPRESSOR BLEED TUBES COMP ROTOR 1 & 2 SPOOL COMP ROTOR SIG 3 DISK COMP ROTOR 4-9 SPOOL COMP ROTOR FWD SHAFT COMPRESSOR STATOR DISTRIBUTORS (AUGMENT) ENGINE ACCY GEARBOX EXHAUST NOZZLE DUCT EXHAUST NOZZLE LINER EXHAUST NOZZLE SHROUD

EXHST NOZZLE EXT DUCT EXHST NOZZLE EXT LINER EXHST NOZZLE OUTER FLAP EXHST NOZZLE PRIM FLAP EXHST NOZZLE PRIM SEAL EXHST NOZZLE DIVERG FLAP EXHST NOZZLE DIVERG SEAL EXHST NOZZLE ACTUAT RING EX/NOZ ACTUAT LINK & BRAC FAN BLADES & VANES FAN FRAME FAN IGV FLAP FAN OUTER DUCT FAN ROTOR STG1/FWD SHFT FAN ROTOR STG 2 DISK FAN ROTOR STG 3 DISK FAN ROTOR AFT SHAFT FAN STATOR CASE/FWD MNT FAN STATOR VANE FRONT FRAME HPT BLADE/VANES HPT ROTOR/FWD/AFT SHAFT HPT STATOR/NOZZLE IDG PIPING LPT BLADE/DOVETAIL LPT ROTOR/SHAFT LPT STATOR/VANES TURBINE FRAME

Valves, Fan and Core Spraybar Valves, Local Distribution Control, Fuel, Hydromechanical (MEC) Control, Eletronic (AFTC) Control, Augmentor Fuel Pump, Augmentor Fuel Pump, Hydraulic Pump, Lube/Scavenge Pump, Main Fuel Pump, Total Fuel Boost Valve, De-icing Core Stator Actuator Variable Stator Vane Feedback Cable Engine Monitoring System Processor Gearboxes Sensor, Flame Ignitor, Main Ignitor, Augmentor Augmentor Filter Electrical Harness Alternator

ON

Table 6.1 Major Components and Subsystems Requiring

Qualification

Actuator, Exhaust Nozzle (AB) Actuator, Fan IGV Cooler, Lube Oil/Fuel Cooler, Hydraulic Oil/Lube Oil Cooler, Lube Oil/Air- (IDG) Detector, Turbine Met Temp (Pyrometer) Exciter, Ignition Sensor, Fan Discharge Temperature (T25) Sensor, Fan Speed Sensor, Inlet Temperature (T2) Transducer, Exhaust Nozzle(AB) Feedback Tank, Lube & Hydraulic Alrcraft/Eng Interface Mounts, thrust Fitting and Links Filter, IDG Oi! Line, Fuel & Motive Flow Pump, Supply Ejector, Air/Oil (IDG) Cooler Regulator, Ejector (9th Stg Bleed Air) Flowmeter, Fuel

1. Load factors and angular velocities and accelerations should be taken at or about the C.C. of the engine.

FLIGHT (0 to Max Thrust) '&'- ±6RAD/SEC2i ** = 0 >

i =o

2. Side load factors (S.L.) set to either side.

) *>FORE

= ±2 RAD/SEC S.L. = 4.0

3. 8 and 8 are pitching velocity and acceleration,

Applicable to complete crosshatched area.

4. $ and \f are yawing velocity and accleration.

Applicable to complete rectangle from 7 UP to 10 DOWN

6. Fore loads occur during arrested landing

5. Down loads occur during pull out.

9 = 0 \ S.L. = 1.5 f

DOWN LANDING (0 to Max Thrust)

AFT«

• FORE

S.L. = 2.0 ^ =0 8= +14Rad/SEC2 £ = ±6RAD/SEC2 CATAPULT (Max Thrust)

UP 1.2 --1

AFT- intermediate power or above, and back to idle. A CIC is defined as cruise power setting to intermediate power or above, and back to cruise power setting. For engine life requirements, TAC = LCF + 0.25 (FTC) 4- 0.025 (CIC) An example of how a mission cycle is converted to an ASMET cycle is shown in Figures 6.15 and 6.16.

6-14

There are many ways of converting from a mission cycle to an accelerated cycle depending on the objectives. For example, if the objective is to demonstrate rotor durability you would want to keep all idle to intermediate excursions and minimize max A/B excursions while if the durability of the exhaust nozzle were to be demonstrated, you would include all max A/B excursions while minimizing idle to intermediate transients. The endurance test may be conducted at various inlet pressures and temperatures, horsepower extraction and fuel temperatures for selected number of cycles. Customer bleed and the de-icing system may be required for a selected number of cycles. As part of the qualification endurance test, a high cycle fatigue (HCF) test is usually included to assure that the tubes and pipes don't have detrimental resonances in the engine operating range. The HCF, known as the stair-step bodie, is typically run in two parts, the up-leg and the down-leg. The up-leg is normally done at the onset of the endurance test and the down-leg at the end. An example of an HCF cycle is shown in Figures 6.17 and 6.18 for the up-leg and down-leg respectively. The HCF cycles consist of a sequence of rotational speeds of one hour duration ranging from idle to maximum speed in increments of 200 rpm. The down-leg is offset by 100 rpm from the up-leg so as to cover every 100 rpm from idle to the maximum operating speed. This sequence will provide adequate data to evaluate the tube and piping characteristics and identify potential problems. Operability Evaluation The operability evaluation is normally conducted in the altitude qualification test to evaluate the effects on transient operation and steadystate performance.

ENGINE QUALIFICATION AND CERTIFICATION -

CO _j

z z r "^ o

<

C9 Z

£S *£ o—

o0 . * u. u

U D o

o ~-

C9Z 2— — X

X

a a>

2a D

UJOJ

OTQ

a

CQ CL

a. a a

CM

uo

H W

U>W

r-

o a>

ft5 mo UJ

z— — X

CCtJ

(9

xcc

tDt-«

<

r> Li.

oo o a> tow —w

CD

N

a.

?

i Ul

o z

f

ice

«>: »x

OJ

,.

to

x to r-

ft u

u. u oo o «-

a a CM

X (O



.

> 0o

Figure 7.11 Speed Effects On Operating Contact Angles 7-8

BEARINGS AND SEALS

This of course only holds true for a bearing taking pure thrust and no radial load so that all balls are equally loaded. At high speed the bearing on the right has significant centrifugal (C.F.) load vectors on each ball. This causes unequal inner and outer contact angles in order to balance the extra radial ball load. The outer race contact angle must get smaller and the inner will correspondingly increase because the total bearing clearance is assumed unchanged. The ball rolling axis will align primarily at right angles to a line normal to the outer race contact and thus increased sliding and heat generation must occur at the inner race because of it's physical displacment from the roll circumference. At high contact angles and high speeds the sliding and heat generation can degrade the lubricant films in the contacts and cause bearing damage. This effect can be reduced by reducing the design contact angle but this of course reduces fatigue life as previously shown. Making raceway curvatures larger (less conforming) will also reduce heat generation at the contacts but this also involves a fatigue life penalty. It will also be noted that during one cage revolution the roll axis of the ball travels in a cone shaped trajectory with the apex of the cone at the centerline of bearing rotation.

"TSd

FA

Thus during one revolution the roll axis must swing through a total angle of 2a. This requires a gyroscopic moment to achieve and the moment must be provided by friction forces at the raceway contacts. At very high speeds the available forces may not be sufficient and gyroscopic spinning will occur at the contacts. This in some cases may cause distress due to additional heat generation and lubricant film breakdown. Gyroscopic spinning can be reduced by reducing the contact angle but as previously stated, a fatigue life reduction results. A further consideration is maximum ball excursion occurring in the bearing. Figure 7.12 shows a bearing with combined radial and thrust loading CO C\J

o o

CD

CO

I

8-2

SECONDARY SYSTEMS

m

to

e CO

b c o o CO

U_l C-3

1 J= tTi



o o CO

CD

o

^J

c .o

* *

3 J 4 6 7

©

©

o.*:i oo

e.»Oo i.W oc

«o«&££fc] „jyuaa @

]«*. CCflMICOul

IfVIMIHlMGt

•-

»AM t O O l l I t

( U K HA* 0 1 NdlUaUAMOMAM

Figure 8.15 Lube System Schematic

Lube Tank The functions of the lube tank are to store oil, deaerate scavenge oil, accommodate inverted operation (if required), vent excess air, provide adequate supply pump inlet pressure, and provide for oil sampling. Lube tank size is determined by estimated engine oil consumption, system gulping (oii quantity in transient during operation), all attitude reserve, and expansion space for oil thermal excursions. Oil tank shape is frequently determined by available space, particularly in military applications. A good example is the J79 tank shown in Figure 8.16. That same figure shows an internal schematic of the tank giving an indication of the complexity of tanks for engines requiring inverted flight. All tanks have a tank pressurizing valve which builds up tank pressure so a positive pressure will always be maintained on the supply pump element to prevent cavitation even at high altitudes. The source for this pressure is air that is returned to the tank with the scavenge oil. The air is routed through a deaerator inside the tank then vented back to the gearbox or overboard through an air/oil separator. Depending on the application, tanks can have manual oil level indicators or electrical cockpit indication. All tanks have remote fill/overfill ports for servicing with a lube cart.

SECONDARY SYSTEMS

Lube Pump The lube and scavenge pump elements are usually housed in the same casting. A typical cross section is shown in Figure 8.17. Pumping elements can be vane, gear or gerotor (internal gear) type elements. Each type is shown in Figure 8.18. Once the engine requirements have been established, the pump specification can be issued and the vendor competition and selection completed. Then the development testing and engine qualification testing follows. The pump body is sometimes used to provide other functions such as chip detector mounting, inlet screen mounting, cold start bypass relief valve, filter mounting, filter bypass'mounting, filter service shutoff valves, or core speed sensor drive/ mounting. Lube Pipe Lines and Jets Lube supply lines are sized to deliver required oil flowrates to the sumps at a maximum velocity of 10 feet per second. This assures that the majority of the system pressure drop will take place at the lube jets. The use of stainless steel tubing is required. Supply lines are routed near the bottom of the engine to minimize effects of soakback temperature increase after shutdown. Lube jet minimum diameter is limited to 0.025 inches to minimize jet plugging.

8-17

Sump Vent Check Valve Scavenge Return Overfill Port Tank Vent Outlet Port(CSD) Lube Supply Port

Hydraulic Supply Lube Sen/ice Port

J79 OIL TANK SCHEMATIC

© ^ ^ ©-at

•S/Sfr

1 S U M P VENT PORT 7 SUMP VENT CHECK VALVE 3 DIVE ANO VENT * CLIMB AND V E N t S R E i i e * VALVE B MASTER VALVE

7 SCAVENGE O i l HETUHN B F ILL PORT (ALTERNATE) 9 OVEHFIIL PORT 10 CSO FLEXIBLE PICKUP M OEAERATOR 11 I N V E R T E O A D V E N T

I] 1> IB IB \7

H T O H A U L C COMPARTMENT AIR VENT H V O f l A U l I C FLEXIBLE PICKUP GRAVITY VALVE TANK VENT POUT LUBE SUPPLY PORT

AHII

I A A W

IB L U 6 E COMPARTUEHT OftAIN A N D FILL P O T T 19 HYDRAULIC SUPPLY PORT

M cso SUPPLY PORT 21 HYORUALIC COMPARTMENT DRAIN

Figure 8.16 J79 Oil Tank

8-18

SECONDARY SYSTEMS

~*

1 1

SCAVENGE

/ C O M M O N SCAVENGE DISCHARGE

/

" D " SUMP SCAVENGE INLET

I

LUBE SUPPLY DISCHARGE ^_ A G B SCAVENGE INLET '-^fy~~~^y-T~LUBE SUPPLY INLET

LUBE

INLET PRESSURE T G B SCAVENGE INLET " B " SUMP SCAVENGE INLET

42.2MB.. Freemake Video Converter - Convert video to AVI, MP4, WMV, MKV, 3GP, DVD, MP3, iPad, iPhone, PSP, Android phones. Video to MP3 with.... 3Crack[Apowersoft. ... 4Apowersoft Video Converter Studio ... v4.8.2()56.1MB ... Freemake Video Converter v4.1.10.252.. Freemake Video Converter v4.1.10.252 Setup + Crack ~ [APKGOD], 6, 0, Jun. 1st &#39;19, 52.7 MB6, apkgod VueScan Pro v9.6.39 Setup + Crack (x64/x86).... Freemake Video Converter v4.1.10.252 Multilingual-P2P. Posted on ... Fit the output file size to any limit (e.g. 700 Mb, 1.4 Gb, 4.7 Gb).

freemake video converter ... Freemake Video Converter v4.1.10.517. 30.82 MB 2020-02-15 ... 3CrackPatch.. Freemake Video Converter v4.1.10.491 + Crack ~ [FileRiver]1, 21, 2, Dec. 27th &#39;19, 51.8 MB21, apkgod. Freemake Video Converter Gold v4.1.9.77 Final +.... Freemake Video Converter Convert video to AVI, MP4, WMV, MKV, 3GP, DVD, MP3, iPad, iPhone, ... Freemake Video Converter v4.1.10.252 + Crack ^g^^^ I^^^^^^^^H I i mmmmmmmmmmmmmmmmm I^^^^^^^^^B Format Factory 4.6.0.2 License Key  - Crack Key For U _ .J

i• ^

i

-

r

-I j_ I \\

-r_ 4 - - 4i

I 1

1 1

19.3MB . crackKeygen.exe8HD Video Converter . Freemake Video Converter v4.1.10.252 0

1 .1

I I

A

i

. -j 1 - - -_ 1X ™_ _Hi

1i

1 .2

.22

PT/PT

.24

,--iq-

1 1 i i

50,

3ol--

loo,

n OPEN CLEARANCE OUGHT CLEARANCE

.28

AB5. FLOW ANGLE

SINGLE-STAGE HPT EXHAUST PRESSURE AND FLOW ANGLE DISTRIBUTIONS FOR TWO LEVELS OF T I P CLEARANCE

Figure 5.21 Single-stage HPT Exhaust Pressure and Flow Angle Distributions for Two Levels of Tip Clearance

CC DR

X lC0

r> < 3

tUJ

o _l u.

_J

H 10 O < X X UJ

CC

z o oCC t-

# u z

01

o

tCO

o m o u u X

• Figure 5.36 Static CDP and FOS Seals 5-36

TURBINES

HP TURBINE BLADE

AFT AIR SEAL

COW-RE SSOR DISCHARGE PRESSURE SEAL ECDPJ

Figure 5.37 HP Turbine Motor

Compressor Discharge Seal (CDP) Disk This component is sized to provide a seal for compressor discharge air as discussed earlier. A disk is required to support the rotating seal head and to keep the concentrated stresses in the bolted flange to levels low enough to avoid low cycle fatigue. Another factor in the size and proportion of this disk is to keep critical rotor vibratory frequencies out of the range of operating speeds. Forward Shaft The purpose of this shaft is to provide a torque and axial load carrying structural member between the HP disk and HP compressor. Again this component is proportioned to avoid critical rotor vibratory modes and low cycle fatigue. Forward Outer Seal (FOS) Disk and Retainer The rotor disk is relatively complex as disks go. As with the CDP seal disk, it provides support for the second (outer) cavity Format Factory 4.6.0.2 License Key - Crack Key For U head, sometimes called the "four tooth seal". Cantilevered off of the seal head is a flexible arm supporting the forward blade retainer. This structure provides assurance that the HP blades will not move forward in the flowpath. The retainer portion of this

TURBINES

structure is rabbeted to me HP disk. Rabbeted construction has three basic purposes. It provides additional radial support as the FOS arm could not withstand bending stresses associated with centrifugal loads. Secondly, this construction assures that rotor parts do not shift under heavy centrifugal loads. Shining would destroy the impeccable balancing of these high speed components. Imbalance associated with shifting of rotor parts would cause heavy one per rev vibratory loads, generate excessive seal leakages, and could generate high cycle fatigue. Finally, rabbetted construction provides a seal for the blade cooling circuit. Blade cooling air would be dumped into the FOS cavity without this seal (see Figure 5.38). Besides being a performance loss, without the cooling air the HP blade would overheat and rupture. The web of the FOS disk has holes placed slightly below the inducer center line. These holes permit the inducer to pressurize the blade cooling circuit. Remember the inducer accelerates the blade cooling air to meet these rotating holes and thereby expanding (cooling) the air. Format Factory 4.6.0.2 License Key - Crack Key For U the air is now moving at wheel speed it can get on board the rotor without heat being generated through windage. This is a double benefit: less drag on the rotor (more efficiency) and increased cooing capacity for the

5-37

Figure 5.38 Secondary Air Flow HP blade. Placing these holes below the center line of the inducer helps keep dust out of the blade cooling circuit. The air can turn the corner and the dust if it is heavy enough will be centrifuged back to the flowpath. Radially, inboard of the inducer holes, is a two tooth seal which prevents blade cooling air from leaking into the CDP cavity or hot CDP cavity air from leaking into the blade cooling circuit as a function of pressure gradients. The FOS disk bolts lo the HP disk and forward shaft. The complex shape of the FOS disk can result in unnecessary thermo-mechanical bending stresses. When these features are carefully positioned with respect to each other these stresses can be minimized. This is called "stacking the disk". A full understanding of the time varying thermal gradients and mechanical loading is the key to successfully "stacking the disk". HP Disk This component supports the blade dead load (over 2 million pounds), along with the radial rabbet loads from the forward and aft retainers. The HP blades are held by a two tanged fir tree arrangement shown in Figure 5.39. The large radial blade load and disk mechanical and thermal stress result in a very large bore

5-38

hoop stress. This stress is reduced by increasing the bore width (see Figure 5.37). Disk flange arms facilitate bolting other shafts and disks to the HP disk. Again this disk is proportioned to avoid critical vibratory modes in the engine operating speed range. Aft Retainer This component is rabbeted to the HP disk and is charged with providing an axial stop for the HP blade. It also provides sealing for the HP blade cooling circuit. When forward and aft retainers seal properly, the blade cooling circuit is pressurized by the inducer and the hollow HP blade is convectively cooled. Aft Shaft This rotating structural member provides support for the HP rotor as its journal houses a radial bearing. The shaft also meters the bore cooling circuit. Shaft holes permit bore cooling air to leave the HP rotor and pass on to the LP rotor, (see Figure 5.38). As with the CDP and FOS disks, this shaft has seal teeth which interact with the static and rotating low pressure (LP) turbine, thereby sealing HP and LP secondary flows from each other. Again, this component is proportion to avoid critical rotor vibratory modes in the operating speed range.

TURBINES

80 DOVETAILS

72 DOVETAILS

CFM56-5

CFM56-3

Figure 5.39 Comparison Between CFM56-3 and CFM56-5 Dovetails

HP Blade The HP turbine has eighty hollow cast turbine blades. The internal passages have been designed to provide for an efficiently cooled turbine airfoil. These circuits are shown in Figure 5.40. Convection is the prime heat transfer mechanism. After the cooling air speeds through the blade passages, it is returned to the flow path gas stream through small holes in the airfoil and tip cap.

The blade, like the turbine disk, must be stacked carefully to reduce bending stresses caused by gas path pressures and centrifugal loads. The aggressive thermal environment can even cause tough blade alloys to oxidize or corrode away so the blade receives a CODEP coating to protect parent blade metal.

THE MECHANICAL DESIGN PROCESS

Air exiting the airfoil forms a thin film of air around the blade to minimize disruption to gas streamlines and to shield the external blade surface from hot flowpath gases. Thin trailing edges and gas path stagnation at blade leading edges cause these locations to be the hottest. Airfoil holes are usually placed in these locations with design intent being to minimize airfoil thermal gradients.

The design of HP turbines is a challenging problem, that has very little margin for error. Complex thermomechanical stresses generated in Malwarebytes Anti Malware 3.5.1 Crack + Premium Key Download operation can produce failure in meeting design goals and sometimes actual engine disfunction. The road to a successful design is attention to detail. Over looked items usually lead to serious problems.

There are three functions for the air exiting the blade tip. The first function is to cool the tip cap, the second is to purge internal blade circuitry and reduce internal pressure loss, and thirdly it purges the blade of dust. If dust accumulates in blade cooling passages or blades cooling holes the result will be a hot blade and rupture/fatigue.

While the dream of every turbine designer is a clean sheet of paper, the reality is a nightmare of deciding where to start. Then, there are seemingly endless iterations of refinement and discovering how to compromise the conflicting design goals and constraints of weight, performance, and durability.

TURBINES

5-39

TIPCAP PURGE HOLES

CAVITY I CROSSOVER HOLES [IMPINGEMENT COOLING)

LEAD EDGE COOLING H X E S

PITCH SECTION

TRAIL EDGE COOLING HOLES

3 CIRCUIT HPT BLADE COOLING Figure 5.40 Cooling Circuitry for CFM56-5 HPT Blade

Whiie flowpath hardware (nozzles and blades), as well as stator and rotor hardware each have their own particular design problems. They also have similarities. To illustrate the process let us focus on the rotor to get a sense of how a mechanical design evolves. An overview of design and development includes the following activities: Basic design concept Preliminary sizing Engineering drawings Working the details Component and engine testing, and Final certification or qualification analyses All designs rely on four critical items, teamwork, experience, demonstration, and verification. A turbine mechanical designer is a conductor of a large orchestra composed of specialist in areas like: Cycles and performance Aerodynamics Secondary flow

5^0

Heat transfer Materials application Drafting Testing Computer methods Stress and vibration analysis, and Manufacturing The designer must meld the design inputs and constraints from each of these areas into a harmonious result. Experience is found in a variety of places technical specialists, other mechanical designers, GE design practices, reviews held by the Chief Engineer's office, and individual designer's own personal reservoir of knowledge and insight. No matter how strong the designer's knowledge and insight is, it will be insufficient to producing a viable design. Experience in applying engineering principles will result in a sound design concept, only when all resources are utilized in a detailed fashion. A design concept can be refined and verified through repeated modeling and analysis. This is an integral portion of design verification. Application of design analysis

TURBINES

methods tempered by lessons learned is also a powerful cost effective method for accumulating and applying actual design experience. Final demonstration of all designs is achieved through component and engine testing. The engine always processes a level of wisdom that humbles the best thought out designs. The Basic Design Systems engineering has prime responsibility MAGIX Movie Edit Pro 2021 20.0.1.73 Crack With Product Key Free the total engine system meeting its requirements. Their relationships with cycles, performance, aerodynamics, and mechanical design specialists gives broad form to the engine and its requirements. Key items here are: Number of turbine stages Flowpath dimensions Aerodynamic shapes for the nozzle and blade Numbers of nozzles and blades Engine speed and thermodynamic parameters, and Locations of frames and bearings Once the size and performance parameters of the HP turbine are established by systems engineering the mechanical design process can begin.

Preliminary Design The core engine rotor structure is usually supported at its extremes by bearings, one of which is an axial thrust bearing. (Figure 5.41). For the CFM56/F101 family of engines the forward bearing supporting the HPC rotor, is the axial thrust bearing. This bearing supports the net axial load on the core rotor. The sources of this load are HPC and HPT blade loading and internal pressures acting on the rotor structure. This bearing is also the reference point for axial growth of the rotor, an important factor in establishing rotor seal design. With the inputs provided by Systems Engineering, the mechanical designer begins discussions with Systems, Secondary Flow, Heat Transfer, Bearings Design and Performance specialists to define the secondary flow aspects of the turbine structure. Fundamental engine issues are being settled here. Major issues include: How much bleed air will be used to cool and purge each rotor-stator cavity and the HP turbine blade? Where will the bleed air be extracted? How will the bleed air be reintroduced to the engine flowpath?

STAGE 1-2 SPOOL

FWD. SHAFT

COMPRESSOR RADIAL BEARING

THRUST BEARING

Figure 5.41 CFM56-3 High Pressure Rotor Bearing Support

TURBINES

5-41

Given the above, what is the impact on engine performance?

is generating engineering drawings to support technical work.

How can engine performance be improved?

An early milestone is the material release drawing for forgings and castings. Lead time for these items can be from six months for forgings to one year for castings and must be allowed to achieve the overall engine schedule. Experience, intuition, and engineering estimates are used to provide detail component definition for fillets, hole size, and other features which can result in significant level of concentrated stress.

What will be the diameter of the compressor discharge and the pressure balance (forward outer) seals? Is the resultant axial load within the bounds adequate for bearing life? Does this secondary flow system present insurmountable mechanical problems? Resolving these fundamental issues requires creative thought, evaluation of many alternative concepts through an iterative process, and then arriving at a design based on mutual compromise. During this process, estimates of internal cavity flows, pressures and temperature are made by secondary flow specialists. The heat transfer specialists estimate metal temperature distributions for generic acceleration from steady state idle to steady state takeoff and a deceleration back to idle. With these estimates and engine speed (a thermodynamic parameter) the mechanical designer begins to size the various turbine rotor disks. Disks are usually required to support rotating seal heads and large dead loads like those created by the rotating turbine blades. The designer must assure that each disk has adequate overspeed margin to its burst speed, and then estimate radial and axial deflections for rotating seal heads as well as the HP turbine blade/ These displacements when compared to deflection estimates made by those designers responsible for the HP stator provide insight to potential seal clearances. These clearances usually require the secondary flow analysis to be repeated, and so on until the rotorstator interaction is understood and secondary flow and performance constraints MAGIX Movie Edit Pro 2021 20.0.1.73 Crack With Product Key Free satisfactorily met. This process requires many months of work until the final design concept is worked out. Engineering Drawings During this early period the mechanical designer will work out associated details relative to sizing shafts and bolting issues. In sizing each HP turbine component the designer must also assure that adequate materials are selected for each component, operating bulk temperature and stress distributions are kept to levels consistent with creep, rupture, and low cycle fatigue life requirements, disk and rotor vibratory modes are kept outside the engine operation range, and drafting

5-42

After much design work and discussion, involving design and design analysis reviews with the Chief Engineer's office and other experienced mechanical designers, the final engineering drawings are released for each component. Again, timing is important due to manufacturing lead time. A sufficient period is required for tooling, gaging, and machining. Typical process requirements are from three to six months. Working the Details From the release of engineering drawings to the first chips made cutting the initial turbine components, a time period is provided for detailed design analysis of the HP turbine rotor. Early in the engine program. Systems Engineering has defined a design flight cycle or cycles for engine wide application (Figure 5.42 and Table 5.4). This operating cycle is a generic description of revenue service engine usage. It is used by design and component systems specialists to establish the transient and steady state turbine operating environment. Rotor heat transfer specialists construct a transient heat transfer model of the rotor (Figure 5.43). The resulting transient metal thermal distribution shown in Figure 5.44 is for the "end of takeoff mission point. This model is based on steady state parameters determined by secondary flow specialists for the design flight cycle (Figure 5.45). Mechanical Design specialists then generate a structural rotor model (Figure 5.46) composed of shells, rings, and two dimensional finite element members. Transient thermal distributions are then input,to the rotor model for approximately 100 time points of interest. This structural model provides basic understanding of bulk stress behavior Format Factory 4.6.0.2 License Key - Crack Key For U the rotor with time (Figure 5.47). To determine levels of concentrated stress around fillets, holes, slot bottoms, and complex non-symmetric rotor features more detailed models are required. Now, two and three dimensional finite element models are generated to fully describe component geometries and predict correct values of concentrated stresses. Shown in Figure 5.48 is the bolted joint where the forward shaft and

TURBINES

]2

J±4

THRUST REVERSE

CD

CO

(X UJ O Q_

TIME Figure 5.42 CFM56-5 Design Flight Cycle Profile and Points

ALTITUDE (FEET)

POWER

MACH NO.

L/M/N/O* DTAMB (°F)

INTERVALS OF TIME (MIN) 1.39

0

.35

-9/11/27/66

1500

.40

-9/11/27/66

10000 20000 30000

.61 .733 .850

10/15/20/40 10/15/20/40 10/15/20/40

20.0

CRUISE

30000

.850

10/15/20/40

19.24

DESCENT

30000 20000 10000

.859 .733 .610

10/15/20/40 10/15/20/40 10/15/20/40

14.0

5000

.3

10/15/20/40

5.0

FIDLE

0

.2

-9/11/27/66

0.11

THRUST REVERSE

0

.2

-9/11/27/66

0.26

GiDLE

0

0

-9/11/27/66

5.0

TAKE-OFF END TAKE-OFF CLIMB

APPROACH

- RELATED TO STD DAY

c o o >

a

n m o

25 o z

SUB-SYSTEMS

COMPONENTS ACCY DRIVE INTERNAL G/B ACCT DRV GRZ BEVEL GEAR ACCT DRV RAD DRV SHAFT ACCY G/B HORIZ SHAFT ACCY G/B HRZ SHAFT HSG A/C ACCESSORY GEARBOX AUGMENTOR MIXER AUGMENTOR DUCT AUGMENTOR FUEL MANIFOLD AUGMENTOR FUEL TUBES AUGMENTOR FUEL VALVE AUGMENTOR IGNITOR AUGMENTOR LINER AUGMENTOR SPRAYBARS COMBUSTOR (IGNITION) COMBUSTOR CASE COMPRESSOR BLADE/VANE COMPRESSOR BLEED TUBES COMP ROTOR 1 & 2 SPOOL COMP ROTOR SIG 3 DISK COMP ROTOR 4-9 SPOOL COMP ROTOR FWD SHAFT COMPRESSOR STATOR DISTRIBUTORS (AUGMENT) ENGINE ACCY GEARBOX EXHAUST NOZZLE DUCT EXHAUST NOZZLE LINER EXHAUST NOZZLE SHROUD

EXHST NOZZLE EXT DUCT EXHST NOZZLE EXT LINER EXHST NOZZLE OUTER FLAP EXHST NOZZLE PRIM FLAP EXHST NOZZLE PRIM SEAL EXHST NOZZLE DIVERG FLAP EXHST NOZZLE DIVERG SEAL EXHST NOZZLE ACTUAT RING EX/NOZ ACTUAT LINK & BRAC FAN BLADES & VANES FAN FRAME FAN IGV FLAP FAN OUTER DUCT FAN ROTOR STG1/FWD SHFT FAN ROTOR STG 2 DISK FAN ROTOR STG 3 DISK FAN ROTOR AFT SHAFT FAN STATOR CASE/FWD MNT FAN STATOR VANE FRONT FRAME HPT BLADE/VANES HPT ROTOR/FWD/AFT SHAFT HPT STATOR/NOZZLE IDG PIPING LPT BLADE/DOVETAIL LPT ROTOR/SHAFT LPT STATOR/VANES TURBINE FRAME

Valves, Fan and Core Spraybar Valves, Local Distribution Control, Fuel, Hydromechanical (MEC) Control, Eletronic (AFTC) Control, Augmentor Fuel Pump, Augmentor Fuel Pump, Hydraulic Pump, Lube/Scavenge Pump, Main Fuel Pump, Total Fuel Boost Valve, De-icing Core Stator Actuator Variable Stator Vane Feedback Cable Engine Monitoring System Processor Gearboxes Sensor, Flame Ignitor, Main Ignitor, Augmentor Augmentor Filter Electrical Harness Alternator

ON

Table 6.1 Major Components and Subsystems Requiring

Qualification

Actuator, Exhaust Nozzle (AB) Actuator, Fan IGV Cooler, Lube Oil/Fuel Cooler, Hydraulic Oil/Lube Oil Cooler, Lube Oil/Air- (IDG) Detector, Turbine Met Temp (Pyrometer) Exciter, Ignition Sensor, Fan Discharge Temperature (T25) Sensor, Fan Speed Sensor, Inlet Temperature (T2) Transducer, Exhaust Nozzle(AB) Feedback Tank, Lube & Hydraulic Alrcraft/Eng Interface Mounts, thrust Fitting and Links Filter, IDG Oi! Line, Fuel & Motive Flow Pump, Supply Ejector, Air/Oil (IDG) Cooler Regulator, Ejector (9th Stg Bleed Air) Flowmeter, Fuel

1. Load factors and angular velocities and accelerations should be taken at or about the C.C. of the engine.

FLIGHT (0 to Max Thrust) '&'- ±6RAD/SEC2i ** = 0 >

i =o

2. Side load factors (S.L.) set to either side.

) *>FORE

= ±2 RAD/SEC S.L. = 4.0

3. 8 and 8 are pitching velocity and acceleration,

Applicable to complete crosshatched area.

4. $ and \f are yawing velocity and accleration.

Applicable to complete rectangle from 7 UP to 10 DOWN

6. Fore loads occur during arrested landing

5. Down loads occur during pull out.

9 = 0 \ S.L. = 1.5 f

DOWN LANDING (0 to Max Thrust)

AFT«

• FORE

S.L. = 2.0 ^ =0 8= +14Rad/SEC2 £ = ±6RAD/SEC2 CATAPULT (Max Thrust)

UP 1.2 --1

AFT- intermediate power or above, and back to idle. A CIC is defined as cruise power setting to intermediate power or above, and back to cruise power setting. For engine life requirements, TAC = LCF + 0.25 (FTC) 4- 0.025 (CIC) An example of how a mission cycle is converted to an ASMET cycle is shown in Figures 6.15 and 6.16.

6-14

There are many ways of converting from Format Factory 4.6.0.2 License Key - Crack Key For U mission cycle to an accelerated cycle depending on the objectives. For example, if the objective is to demonstrate rotor durability you would want to keep all idle to intermediate excursions and minimize max A/B excursions while if the durability of the exhaust nozzle were to be demonstrated, you would include all max A/B excursions while minimizing idle to intermediate transients. The endurance test may be conducted at various inlet pressures and temperatures, horsepower extraction and fuel temperatures for selected number of cycles. Customer bleed and the de-icing system may be required for a selected number of cycles. As part of the qualification endurance test, a high cycle fatigue (HCF) test is usually included to assure that the tubes and pipes don't have detrimental resonances in the engine operating range. The HCF, known as the stair-step bodie, is typically run in two parts, the up-leg and the down-leg. The up-leg is normally done at the onset of the endurance test and the down-leg at the end. An example of an HCF cycle is shown in Figures 6.17 and 6.18 for the up-leg and down-leg respectively. The HCF cycles consist of a sequence of rotational speeds of one hour duration ranging from idle to maximum speed in increments of 200 rpm. The down-leg is offset by 100 rpm from the up-leg so as to cover every 100 rpm from idle to the maximum operating speed. This sequence will provide adequate data to evaluate the tube and piping characteristics and identify potential problems. Operability Evaluation The operability evaluation is normally conducted in the altitude qualification test to evaluate the effects on transient operation and steadystate performance.

ENGINE QUALIFICATION AND CERTIFICATION -

CO _j

z z r "^ o

<

C9 Z

£S *£ o—

o0. * u. u

U D o

o ~-

C9Z 2— — X

X

a a>

2a D

UJOJ

OTQ

a

CQ CL

a. a a

CM

uo

H W

U>W

r-

o a>

ft5 mo UJ

z— — X

CCtJ

(9

xcc

tDt-«

<

r> Li.

oo o a> tow —w

CD

N

a.

?

i Ul

o z

f

ice

«>: »x

OJ

.

to

x to r-

ft u

u. u oo o «-

a a CM

X (O



.

> 0o

Figure 7.11 Speed Effects On Operating Contact Angles 7-8

BEARINGS AND SEALS

This of course only holds true for a bearing taking pure thrust and no radial load so that all balls are equally loaded. At high speed the bearing on the right has significant centrifugal (C.F.) load vectors on each ball. This causes unequal inner and outer contact angles in order to balance the extra radial ball load. The outer race contact angle must get smaller and the inner will correspondingly increase because the total bearing clearance is assumed unchanged. The ball rolling axis will align primarily at right angles to a line normal to the outer race contact and thus increased sliding and heat generation must occur at the inner race because of it's physical displacment from the roll circumference. At high contact angles and high speeds the sliding and heat generation can degrade the lubricant films in the contacts and cause bearing damage. This effect can be reduced by reducing the design contact angle but this of course reduces fatigue life as previously shown. Making raceway curvatures larger (less conforming) will also reduce heat generation at the contacts but this also involves a fatigue life penalty. It will also be noted that during one cage revolution the roll axis of the ball travels in a cone shaped trajectory with the apex of the cone at the centerline of bearing rotation.

"TSd

FA

Thus during Format Factory 4.6.0.2 License Key - Crack Key For U revolution the roll axis must swing through a total angle of 2a. This requires a gyroscopic moment to achieve and the moment must be provided by friction forces at the raceway contacts. At very high speeds the available forces may not be sufficient and gyroscopic spinning will occur at the contacts. This in some cases may cause distress due to additional heat generation and lubricant film breakdown. Gyroscopic spinning can be reduced by reducing the contact angle but as previously stated, a fatigue life reduction results. A further consideration is maximum ball excursion occurring in the bearing. Figure 7.12 shows a bearing with combined radial and thrust loading CO C\J

o o

CD

CO

I

8-2

SECONDARY SYSTEMS

m

to

e CO

b c o o CO

U_l C-3

1 J= tTi



o o CO

CD

o

^J

c .o

* *

3 J 4 6 7

©

©

o.*:i oo

e.»Oo i.W oc

«o«&££fc] „jyuaa @

]«*. CCflMICOul

IfVIMIHlMGt

•-

»AM t O O l l I t

( U K HA* 0 1 NdlUaUAMOMAM

Figure 8.15 Lube System Schematic

Lube Tank The functions of the lube tank are to store oil, deaerate scavenge oil, accommodate inverted operation (if required), vent excess air, provide adequate supply pump inlet pressure, and provide for oil sampling. Lube tank size is determined by estimated engine oil consumption, system gulping (oii quantity in transient during operation), all attitude reserve, and expansion space for oil thermal excursions. Oil tank shape is frequently determined by available space, particularly in military applications. A good example is the J79 tank shown in Figure 8.16. That same figure shows an internal schematic of the tank giving an indication of the complexity of tanks for engines requiring inverted flight. All tanks have a tank pressurizing valve which builds up tank pressure so a positive pressure will always be maintained on the supply pump element to prevent cavitation even at high altitudes. The source for this pressure is air that is returned to the tank with the scavenge oil. The air is routed through a deaerator inside the tank then vented back to the gearbox or overboard through an air/oil separator. Depending on the application, tanks can have manual oil level indicators or electrical cockpit indication. All tanks have remote fill/overfill ports for servicing with a lube cart.

SECONDARY SYSTEMS

Lube Pump Format Factory 4.6.0.2 License Key - Crack Key For U lube and scavenge pump elements are usually housed in the same casting. A typical cross section is shown in Figure 8.17. Pumping elements can be vane, gear or gerotor (internal gear) type elements. Each type is shown in Figure 8.18. Once the engine requirements have been established, the pump specification can be issued and the vendor competition and selection completed. Then the development testing and engine qualification testing follows. The pump body is sometimes used to provide other functions such as chip detector mounting, inlet screen mounting, cold start bypass relief valve, filter mounting, filter bypass'mounting, filter service shutoff valves, or core speed sensor drive/ mounting. Lube Pipe Lines and Jets Lube supply lines are sized to deliver required oil flowrates to the sumps at a maximum velocity of 10 feet per second. This assures that the majority of the system pressure drop will take place at the lube jets. The use of stainless steel tubing is required. Supply lines are routed near the bottom of the engine to minimize effects of soakback temperature increase after shutdown. Lube jet minimum diameter is limited to 0.025 inches to minimize jet plugging.

8-17

Sump Vent Check Valve Scavenge Return Overfill Port Tank Vent Outlet Port(CSD) Lube Supply Port

Hydraulic Supply Lube Sen/ice Port

J79 OIL TANK SCHEMATIC

© ^ ^ ©-at

•S/Sfr

1 S U M P VENT PORT 7 SUMP VENT CHECK VALVE 3 DIVE ANO VENT * CLIMB AND V E N t S R E i i e * VALVE B MASTER VALVE

7 SCAVENGE O i l HETUHN B F ILL PORT (ALTERNATE) 9 OVEHFIIL PORT 10 CSO FLEXIBLE PICKUP M OEAERATOR 11 I N V E R T E O A D V E N T

I] 1> IB IB \7

H T O H A U L C COMPARTMENT AIR VENT H V O f l A U l I C FLEXIBLE PICKUP GRAVITY VALVE TANK VENT POUT LUBE SUPPLY PORT

AHII

I A A W

IB L U 6 E COMPARTUEHT OftAIN A N D FILL P O T T 19 HYDRAULIC SUPPLY PORT

M cso SUPPLY PORT 21 HYORUALIC COMPARTMENT DRAIN

Figure 8.16 J79 Oil Tank

8-18

SECONDARY SYSTEMS

~*

1 1

SCAVENGE

/ C O M M O N SCAVENGE DISCHARGE

/

" D " SUMP SCAVENGE INLET

I

LUBE SUPPLY DISCHARGE ^_ A G B SCAVENGE INLET '-^fy~~~^y-T~LUBE SUPPLY INLET

LUBE

INLET PRESSURE T G B SCAVENGE INLET " B " SUMP SCAVENGE INLET

52 MB. 14 min ago 16 comments. seretnow.me converter freemake crack mb patch. Faasoft Video Converter v5.4.23.6956 ^^^^ is the ratio of the modulus of toughness of the case material to a known test material. Calculation of actual casing thickness takes into account the hoop load carrying portion of any ring stiffeners as well as the thickness of the shell skin itself.

Casing walls are also expected to prevent the loss of failed rotor blades and the resulting secondary debris. Such fragments have a great deal of kinetic energy and if released they could (and have) penetrate the aircraft

Buckling There are two types of instability against which the casing must be designed, general instability in which the entire shell structure collapses, and panel instability in which the local sections of shell skin

FANS AND COMPRESSORS

3-63

1.0 TITANIUH

INCREASED CONTAINMENT HARGIN

K- .00215

0. 5 — STEEL

m lit

K-

.00059

n p

T Z

( 1

z

0. 1

t-

in

K

rr

m nt

z

.05

V u

.oa

DECREASED CONTAINMENT MARGIN

1.000

10.000

100.000

IMPACT ENERGY

.000.000

CFT/LBS)

Figure 3.S7 Empirical Containment Boundary

partitioned off by the stiffening rings and/or embossments buckle. The buckling of short cylindrical shells was extensively investigated by NACA (ref NACA TN 3786) in the late 1940s. Empirically derived buckling relationships were developed from which most shells can be correctly sized to provide the desired margins over engine loads. Interaction relationships between the overturning moments and torsional moments were also developed and are as follows (see Figure 3.58): M

/ T V

Mn * \ Tn /

Other Failures The ability of the casing to withstand overpressures without rupture and to operate without leakage under normal pressures requires that substantial attention be paid to the sizing of bolted joints. In addition to leakage, rotor failure also is a major joint design factor, for while the casing is not expected to contain major rotor structural failures, it is expected that flange joints will remain intact during such events.

(3-20)

where MQ and TQ are the values of these loadings that would individually cause general instability. Vibration Casing walls located adjacent to the rotor blade tips are subjected to internal pressure pulses traveling with the rotor blades. In addition, low integral order rotor per rev and bearing passing frequency excitations are transmitted to the casing through its joints with the frames. Both the overall vibratory response of the casing and the response of local sections of the casing must be studied to insure that no interaction will exist with potential sources. This is generally done through

3-64

finite element analysis of the casing and the Campbell div agram approach. * """""

The development of safe and effective rotor shroud materials is also an important issue in casing design. Materials must be sufficiently soft that they do not cause excessive heating or fracture of the blade tips and sufficiently hard that they do not rut or tear under blade rubbing. In the late 1970s, it was discovered that some tip coatings would explode if a sufficient quantity of coating dust was exposed to high temperature and pressure. This is precisely what can happen during a blade out failure during which the shroud coatings are severely rubbed out by the remainder of the rotor blades which are operating under high unbalance conditions. Since that time, combustion testing has been a key ingredient of all new coating development work.

FANS AND COMPRESSORS

T

Figure 3.58 Cai

CASINGS TRADEOFFS AND IMPLICATIONS FOR DESIGN The need for accurate mounting of vanes and accessories and a high degree of concentricity between the outer flowpath surface and the rotor bearing, and the minimization of weight results in considerable emphasis on manufacturing process development and control. The development of low density, very stiff structures is necessary to achieve these goals, and considerable effort has been expended in this direction. As previously mentioned, composite outer rings are beginning to be used to satisfy containment requirements, thus allowing a lighter weight design for the basic casing shell. Nonconventional manufacturing processes such as chemical milling and creep forming/diffusion bonding are also being used or tested. For the longer term, all composite casings are being considered, though issues of cost and dimensional control remain to be resolved.

FAN AND COMPRESSOR VARIABLE VANE ACTUATION SYSTEMS Actuation systems are multi-component mechanisms which translate the electrical or hydromechanica! output

FANS AND COMPRESSORS

Load Schematic of the engine control into mechanical rotation of stator vanes. As shown in Figures 3.59 and 3.60, the components of the system consist of an actuator, a drive or bellcrank mechanism, a unison ring for each variable stage, and a lever arm for each variable vane. In order to take up misalignments resulting from motion of the system and manufacturing tolerance, ball bearing joints are used in some places. Elsewhere, lined journal bearings are used. By allowing the vanes to rotate, the operating efficiency of the compression components can be optimized at several points of the cycle instead of just at the cruise condition. This allows the aero-designer considerable freedom in his work. Typically, a test is conducted in the early stages of engine use in which the stator positions are modified systematically and both stall margin and performance are measured. From this test a "stator schedule" is established which optimizes performance while maintaining acceptable stall margins. This schedule is imbedded in the control logic, as a relationship between rotor corrected speed and actuator position. The function of the actuation system is to translate the motion of the actuator into the correct repositioning of the vanes.

3-65

BELLCRANKS

UNISON RING ACTUATOR

BELLCRANK MECHANISM

CFMS6 (NON-LINEAR) BELLCRANK MECHANISM

SYSTEM SCHEMATIC LEVER ARM UNISON RING ENGAGEMENT PIN

SELECTIVE FIT SPACER ASYMMETRIC TANGS

ASYMMETRIC " D " SCOT WASHER

VARIABLE VANE

TAPERED OR CLOSE TOLERANCE PARALLEL FLAT ("D" SCOT) BEARING THRUST FACE CASING BASE "HIGH BOSS-' VANE/LEVER ARM CONNECTION SYSTEM

"LOW BOSS" VANE/LEVER ARM CONNECTION SYSTEM

Figure 3.59 Actuation System Terminology

3-66

FANS AND COMPRESSORS

PIVOT POINT

CF6 PUMPHANDLE SYSTEM

STAGE 1

STAGE!

STAGES

STAGE 4 • STAGE 5

TF34 TORQUE SHAFT SYSTEM It

FEEDBACK

Format Factory 4.6.0.2 License Key - Crack Key For U Figure 3.60 Linear and Torque Type Belicrank Mechanisms

FANS AND COMPRESSORS

3-67

ACTUATION SYSTEM DESIGN CONSIDERATIONS Normal operating loads on the actuation system are low, requiring only sufficient force to react the airfoil gas loads and to overcome friction in the system. However, stall and other abnormal operating conditions can result in overpressure pulses of up to ten times normal operating values. This results in designs which are very beefy and dovetail nicely with the need to minimize deflections of the system during normal operation since such deflections result in error in the vane positioning. Because there are so many bearing surfaces in the actuation mechanisms, wear is also a major design issue. Most journal bearings and some ball bearing joints are lined with friction reducing materials such as Teflon® in order to extend the life of the mating metal parts. Mechanical vibration of any of the system components can accelerate the normal wear process by orders of magnitude, and hence it must be avoided in the design of components. Vibration stimuli include the vane natural frequencies and blade passing drivers on the vanes, and the engine per rev and bearing passing frequencies transmitted to the system components through the frames and casings to which they are mounted. In addition, flow pulsing in hydraulic actuators or overly sensitive control feedback systems can cause small motions or "dither" cycles of the system which have much the same effect as vibration on the wear of the system. ACTUATION SYSTEM FAILURE MECHANISMS Wear Key to the effective functioning of the actuation system is the minimization of hysteresis, a phenomenon which results in inaccurate positioning of the vanes, which can in turn result in dangerously high blade stresses. Hysteresis is primarily the result of looseness in the fit-up of the actuation system components and mechanical deflection of the components under load. While this can be minimized initially by designing the components to be very stiff and by specifying very tight tolerances on mating features, wear of these surfaces against one another will eventually occur. Most fan and compressor actuation systems contain numerous interfaces in which metal against metal, metal against coated surfaces, or metal against special low friction inserts or bushings is gradually worn due to relative movement of the mating surfaces. It is important to fully test these interfaces and to quantify and document wear observed both in factory and flight testing (there is sometimes a wide variation in wear between the factory and field environments), since these data can have a profound impact on the rate of inspection and refurbishment

3-68

required in service. Undetected failure of these "minor" parts has led, on at least one occasion, to failure of the variable vane, ingestion of vane into the engine, and consequent massive damage to the compression section of the engine. Hammershock For reasons not fully understood, it is possible to encounter strong shock waves in the gas flow which can result in large pressure pulses across the vanes and consequently very large loads on the components of the actuation system. Delayed augmentor lightoff, nozzle instability, and other causes which cannot generally be predicted during the design of the engine can result in pressures ten times higher than those expected in normal operation. In many cases, these loads size the actuation system components and the mounting of the components. Since mis is in the same direction as the stiffness needed to minimize vane hysteresis, it does not usually result in a major weight penalty to the design.

ACTUATION SYSTEM TRADEOFFS AND IMPLICATIONS FOR DESIGN Because of the need for increasingly efficient turbomachinery, it is likely that there will be a need for more, rather than fewer variable stages, on future engines. This will increase die complexity of the systems and increase the need for wear resistant materials that will function effectively at elevated temperatures. In most cases today, wear effectiveness is available, but at a very steep price. The designer must balance issues of initial cost against maintainability cost in selecting these materials. With the advent of the UDF, practical application of blade actuation systems has arrived. The rotating environment greatly complicates the task of the actuation system designer, but performance and weight tradeoffs for a fully variable cycle engine are simply too good for this to stand in the way.

FAN AND COMPRESSOR SYSTEM DESIGN CONSIDERATIONS Containment & Vibratory Weak Link Criteria Rotor mechanical designers must guard against catastrophic failure mechanisms. A common design practice in stationary steam and gas turbine applications set the rotors to release buckets from their dovetail attachments at a predetermined speed prior to rotor disk burst speeds. Thus eliminating the need to design containment casings for heavy rotor disk fragments which would have more

FANS AND COMPRESSORS

kinetic energy. In aircraft engines where engine operation and inlet conditions are significantly more varied than stationary engines the philosophy of having all blades flung from the rotor simultaneously is impractical. Therefore, axial flow compressors and fans are designed to have the casings contain all of the expected damage from ingestion events, such as tire treads and birds, which are certain to occur. All rotor fragments that are released are expected to exit away from the plane either forward or aft, and fragments exiting the engine radially must have little or no kinetic energy left to cause destructive damage to important airframe control lines or fuel tanks and lines. Further, flight weight aircraft engines employ a "weak link" criteria that simply stated is the airfoil shank is stronger than the airfoil, the blade attachment or dovetail is stronger than the blade shank, and the disk attachment is stronger than the blade attachment. This leads the design to have the lightest fragments fail first, if at all. Two benefits come from applying this criteria. First, in the event of a blade failure, less damage will be done from the smaller fragment, and the engine can be designed lighter as there will be lower unbalance loads and less containment capability required. Secondly, with the dovetail stronger than the airfoil the disk can be designed to prevent a "domino" effect where the loss of one blade results in a disk dovetail post failure which would result in the loss of another blade and so on. System Vibration and Balance Modem jet engines are analyzed for overall system vibrations as complicated spring-mass systems. In these models, the interaction of rotor vibratory excitations and static structure flexibilities can be studied to find critical rotor speeds where bearing and engine mount loads are affected. Very often early in the design phase an important speed operating range is found to have a critical crossing. These can sometimes be moved out of the operating range by changing the stiffness of one or more components, In other cases, the engine control is modified to avoid steady-state operation at a known engine resonance speed. The system vibration models of development and production engines cover not only the engine system but also the mount attachments and airframe installations. Manufacturing tolerances prohibit making perfectly balanced and centered parts and thus all high speed turbomachinery is balanced during assembly. Rotors weighing hundreds of pounds are dynamically balanced to obtain residual imbalances of 10 to 25 gm-in, which is equivalent to the weight of a penny displaced just 3 to 7 inches from the axis of rotation. Long flexible shafts and rotor spools require multi-plane balancing. Indeed, multistage non-integral compressor spools are balanced stage by stage. Then assembled together and rebalanced

FANS AND COMPRESSORS

as an assembly. This procedure tends to reduce bending stresses caused by local couples induced by imbalances during rotation. Aside from vibratory excitations caused by rotor imbalances, the axial flow turbomachinery consisting of multiple stages of rotors and stators constitute the basic elements of an acoustic siren. The wide speed excursions and various combinations of rotor airfoil to stator airfoil ratios, present a device capable of generating a wide variation of frequencies. In point of fact, sirens are used in airfoil component testing to excite the airfoils natural frequencies. Thus engine development plans utilize three methods of designing to avoid harmful vibratory resonances: computer model vibratory analysis, component testing verification of the analytically found natural frequencies and mode shapes, finally, engine testing where actual vibratory responses and effects of damping can be studied. Stress Analysis of Rotor Disks Before we begin the analysis of disks we will have a short review of the basic loads and stress equations governing this type of analysis. Modern turbomachinery designers rely heavily on large finite element computer models to aid in a detailed investigation of the design problem. However, the equations and concepts given here still find usefulness in checking the very large and costly computer model, so that the designer knows the answers are correct. A rotor disk is nothing more than a big ring of material and the most significant stresses are the radial and tangential components. As a holdover from the days when barrels were commonplace, the tangential stress component is commonly known as Format Factory 4.6.0.2 License Key - Crack Key For U "hoop" stress. In simplified flat disk theory it is usually shown that the axial stress and the three shear stress components are zero by reasons of symmetry, but this is not the case in complicated geometries with complex loadings. Here we will develop equations that apply to flat disks as well as other shapes, but are limited in scope as they are one dimensional derivations. Consider a ring of material which is pulled radially from the outside diameter with a uniformly distributed load, P, as shown in Figure 3.61. As the ring deflects radially outward under this loading, tangential forces (manifesting as hoop stress) are developed inside the ring to keep the ring in static equilibrium. The tangential force can be found from the free body diagram in Figure 3.61 by summing forces. Here we see that written as

£F y = 0 = \ PsinOr0de - 2 FH0OP

(3-21)

3-69

P C*/IND

P

F HOOP

O/IN)

F HOOP

Figure 3.61 Radial Pull on a Ring

2 F H00P = PrD j sine d6 = Pr 0 [-cos9

J = 2 Pr0 (3-22)

-'• PHOOP -

°HOOP



Hoop stress in a ring can also be obtained from the mechanical load radial growth (actually circumferential strain) as shown below.

Pf„

Pr„

°6 =

Where TBULK ' S t n e volumetric average temperature in the ring, E is Young's Modulus and a is the coefficient of thermal expansion.

ALE For a rotating ring the formula for hoop stress can be found in most handbooks as p w 2 rCG3

This result can also be derived from the load case by substituting p A Format Factory 4.6.0.2 License Key - Crack Key For U for P and TCG for r 0 in the above integral. The thermal hoop stress in an axisymmetric ring at any radius is given by

3-70

— r

E a {T BlJLK

(r + b) 2xr

2T

r

(3-25)

(3-23)

o» =

"THERMAL

2T

if)

(3-24)

Please note that, when working with total radial growths, the thermal component of growth is removed first. This is because uniform thermal growth does not create stresses. Thermal gradients produce stresses! Numerical Example for Rotor Disk Stresses Consider a fan rotor stage whose blade tip diameter is 45 in with a blade tip speed of 1400 feet per second. The blade root is

FANS AND COMPRESSORS

located at the 22 in diameter with the blade shank and dovetail occupying the area from a 22 in diameter to a 18 in diameter, which is therimof the live disk (actual hoop load carrying part). The geometry of the proposed disk is given in Figure 3.62 along with the material data and temperatures. There are 38 blades, each weighing 1.5 lbs, and the additional dead weight (parts that will not carry any of the hoop load) is approximately 20% of the blade weight. The total dead weight composed of the blades, dovetails, blade retainers, disk posts, etc. may be considered to act as a mass whose center of gravity is located at a 24.5 in diameter. You are expected to find the average tangential stress in the disk, the mechanical and thermal hoop stresses in the disk bore and rim, the live disk rim radial stress, the stage weight, and the "upper bound" estimate for the disk burst speed.

lier. First, find the rpm of the disk from the tip speed and tip diameter.

N=

1400 ft/sec z 45 in

xD

60 sec min

12 in ft = 7130 RPM (3-26)

Then, the angular velocity in rads/sec. _

W-

N T

_

-7tin

-=7T = 7130

30

A lot of preliminary calculations are needed to find the loading definitions to fit the equations we developed ear-

r e v

2v rad

:— •

min

1 min 60 sec

rev Format Factory 4.6.0.2 License Key - Crack Key For U 746.7

rad sec (3-27)

TEMP

180°F

A = 6.094 IN

180°F

RCG - 7.233"R

MATERIAL Tl 6-4 120°F 3 P=. 161 */IN E = 16.5 X 109 PS I @ 150°F CC = 4. 6

X 10^ I N / I N °F (© 150°F 120°F 12D°F 11 OOF

Figure 3.62 Live Disk Geometry

FANS AND COMPRESSORS

3-71

Then, the total rim dead load. ° 8 RIM



= 43.7 KSI

^ e AVG '

NWr g

7.233 in 9.0 in = 35.1 KSI

38BLDS ' I 5 lbs - 1.2- p746.7 - ^ - 1 " ^

386.4

-

(3-32)

in

lb m - in lb f - sec^ (3-28)

There is no average tangential thermal stress as the average tangential temperature is the same as the bulk temperature. So, calculating the thermal hoop stresses at the bore and rim from equation 3-24, we have.

FCENT = 1.209 x JOMbf

This is a big number that is typical of most engines. Now let us express it as a load per inch of circumference on the live disk rim.

P =

1-209 x lQMbf TT • 18 in

TD„

= 2j 3g0

*

o&r

— E a (TBULK

TBORE)

'BORE

= 16.5 x 106 PSI (

4 ft x 1 0 - 6

„J

I (150°F - 110T)

Jbf_ in

= 3.0 KSI (3-33)

(3-29) Ok, now we can use our earlier equations and find the average hoop stress by superposing the equations for hoop stress due to rim load and hoop stress due to rotation.

a&

r D ,

=

E a

^TBULK

-

TR1M)

RIM

= 16.5 x [06 PSI (

4 6

'

y

0

"

6

) (150°F - 180°F)

Pt»2 r£G

Pr„

= - 2. 3 KSI

°9AVG —

(3-34) CT

9AVC



21,380 Ibf • 9 in 6.094 in^

The rim radial stress is simply a load per area calculation and can be quickly found by dividing P by the rim width.

.161 lb/in' (746.7 - ^ - • 7.233 i n ^ sec 386.4

Ibm * in Ibf • sec-

21380 W

lb

2.75 in

= 7.8 KSI

(3-30) °swc = 43.7 KSI

(3-35)

Using equation 3-25, we see that the bore and rim mechanical hoop stress can be obtained by radius ratioing if we ignore any differences in radial growth from the bore to the rim. Thus, we have

°e BORE



i"cc °e AVC *

[

BOR£

= 43.7 KSI •

7.233 in 4.0 in = 79.0 KSI (3-31)

3-72

The stage weight is just the sum of the weights of all of the individual pans, and all are known with the exception of the live disk weight which is

WEIGHT = PA2TRCC = (.161 Ib/m1) 6.094 in-(2?r) 7.233 in = 44.61b (3-36) And thus, the stage weight is 68.4 + 44.6 = 113.0 lbs.

FANS AND COMPRESSORS

Finally, a first pass calculation for disk burst that gives the "upper bound" burst speed occurs when the average tangential stress is equal to the ultimate tensile strength (UTS) of the disk material. Noting that the dead load and centrifugal load stresses are proportional to rotor speed squared, the burst speed can be found to be

= N"\/•

UTS °QAVG

= 7130

J

120 KSI 43.7 KSI

11.815 RPM (3-37)

A word about the accuracy of the stresses and burst speed just calculated is warranted. Comparison to a computer model is made in Table 3.6. This model includes the material property variation with temperature and accounts for the strain variation from rim to bore. As shown in the table, all of the stresses calculated for the example agree with the computer model answers within 2 KSI, except for the bore stress which is 13% high (9.2 KSI). The burst speed calculation that would give a maximum "safe" speed just prior to burst failure would be 10 to 15% below that calculated above. Whereas the "true" burst speed would fall somewhere between the two extremes.

mum thickness that can be used to stop the fan blade of the previous example when failure occurs during an overspeed event 20% above the design speed of 7130 rpm. Of course any weight savings found by using a thinner duct is to be calculated. Let us begin by calculating the kinetic energy of the fan blade. Just prior to failure the kinetic energy is given by the rotational kinetic energy of the blade rotating about the engine centerline, and just after failure the blade kinetic energy is the sum of the translations! kinetic energy license key for adguard the rotational kinetic energy of the blade revolving about its own center of gravity. For ease of computing the kinetic energy of the blade just prior to release is used, and the resulting error is slight (less than 10%). Therefore, the blade kinetic energy is ma.- r cc

Eb.=

l.51bm (7130 RPM) 1.2 Eb = 2 32.2

E b - 19.500 ft • lbs

7T \ 24.5 / 1 ft 30 / 2 V 12 in Ibf • ft Ibf • sec= (3-38)

We can see that the required thickness of the casing is

Casing Containment Capability Suppose we are given a ribbed fan casing, shown in Figure 3.63, to analyze for containment capability and asked to find the mini-

U*« = .00215

in -^.19,500 ft - lb = 0. 3 Vft • Ibf (3-39)

DISK STRESS COMPARISON STRESS 0

B A VG °©BORE

°eRiM 0

©T BORE ©T RIM °TtlM ff

HAND CALCULATION 43.7 KSI 79.0 35.1 3.0 -2.3 7.8

COMPUTER 45.0 KSI 69.8 32.9 4.1 -2.4 7.8

Table 3.6

FANS AND COMPRESSORS

3-73

.25 5 PLACES

4

25 • 1 PLACES

-ACES

Notice: Undefined variable: z_bot in /sites/mynewextsetup.us/license-key/format-factory-4602-license-key-crack-key-for-u.php on line 100

Notice: Undefined variable: z_empty in /sites/mynewextsetup.us/license-key/format-factory-4602-license-key-crack-key-for-u.php on line 100

Comments

  1. Hi there, When i got to the site i dont know which one i should download theres like 4 of them like backtrack 5 r1 and r2 r3 i dont get it which one do i download for windows xp?? 32bit??

  2. But if you get the population birth & death conditions correct, and seed opportunistic locations in the field like you're showing, then it's really some sorta special mojo maker. Kinda like birthing lil' creative cell-kids and seeing them propagate their song.

  3. vlc does not look like yours when I open it. I downloaded it today but screen is not black and media button is not on screen

Leave a Reply

Your email address will not be published. Required fields are marked *