ML20093C467
| ML20093C467 | |
| Person / Time | |
|---|---|
| Site: | River Bend  |
| Issue date: | 04/30/1983 |
| From: | SEAWORTHY ENGINE SYSTEMS, INC. |
| To: | |
| Shared Package | |
| ML20093C471 | List: |
| References | |
| 123-01, 123-1, NUDOCS 8401260415 | |
| Download: ML20093C467 (133) | |
Text
. -
P EVALUATION OF THE OPERATIONAL AND MAINTENANCE HISTORY OF, AND RECENT MODIFICATIONS TO, THE MAIN ENGINES IN THE M.V. COLUMBIA 4
SES Report No. 123-01 Prepared For:
Department of Transportation and Public Facilities State of Alaska Division of Marine Highway Systems P.6. Box 1467 Juneau, Alaska 99802 Prepared By:
Seaworthy Engine' Systems, Inc.
36 Main Street Essex, Connecticut 06426 d /40@9 April 1983
1 I
- ERRATA. SHEET -
EVALUATION OF THE OPERATIONAL AND MAINTENANCE HISTORY OF, AND RECENT MODI FICATIONS TO, THE MAIN ENGINES IN THE M.V. COLUMBIA SES REPORT NO. 123-01, APRIL 1983 1.- Add:. Pg 2-2 (after last item in 3.):
- The new C-17 turbochargers must be properly matched to the engine revised performance ratings.
- The turbocharger installation shall include such modifications as necessary to maintain the waste heat recovery systems at present output level.
- 2. Add: Pg 2-2:
Word " Carbon"
- combustion chamber carbon deposits
- 3. Correct: Pg 2-7 (Table 2.2):
Column Predicted By Fuel Rack ( }
Entry should read: 28.3/6200 BHP
- 4. Correct: Pg 2-12 (2nd paragraph, line 11):
Omit "the fact that" where repeated
- 5. Correct: Pg 2-14 (Table 2.4):
Column 385/4680/135.2-Halter Marine Test Entry should read: 390/5521/147
- 6. Correct: Pg 2-16 (1st paragraph, line 13):
Sentence should read: Rating of $284 BHP /385 ERPM
- 7. Correct: Pg 2-18 (lowest curve):
- Entry should read
- 3/25/83 Test Data, BHP By Rack, Figure 2.1, Fuel By Rack (Ref. No. 3)
- 6. Correct: Pg 2-21 (1st paragraph, line 4):
Sentence should read: ........ Appendices F and G:
- 9. Add: Pg 2-24 (Table 2.7)
- Omitted - 'f2,
- Vertical' columns should read: Left Bank' TC Right Bank-j}TC Average-f TC
- 10. Correct: Pg 3-32 (1st paragraph, line 3):
Omit " generating time"
, Sentence should read: Original for the five (5)................ planning records show that
TABLE OF CONTENTS Section Page 1.0 INTRODUCTIr.N AND EXECUTIVE
SUMMARY
1-1 1.1 Background 1-1 1.2 Executive Summary 1-2 2.0 ANALYSIS OF DE-RATED ENGINES AND NEW 2-1 TURBOCHARGERS 2.1 Engine Performance 2-5 2.2 Turbocharger Performance 2-20 2.3 Post Trial Performance 2-28 3.0 HISTORICAL REVIEWS AND ANALYSIS OF 3-1 COMPONENT FAILURES 3.1 Introduction 3-1 3.1.1 Data Sources 3-1 3.1.2 Time Period 3-2 3.1.3 Chronological Methodology 3-2 3.2 Maintenance History Tabulations 3-2 3.3 Summary of Maintenance / Failure History 3-7 3.3.1 Cylinder Heads 3-7 3.3.2 Cylinder Liners 3-11 3.3.3 Pistons 3-15 3.3.4 Master and Link Connecting Rods 3-16 3.3.3 Camshafts 3-20 3.3.6 Main Bearings 3-20 3.3.7 Cylinder Block 3-21 3.3.8 Major Overhauls 3-29 3.3.9 Turbochargers 3-29 3.4 Summary of Findings 3-30 4.0 ESTIMATED POTENTIAL REDUCTION OF COMPONENT 4-1 FAILURES AFTER ENGINE DE-RATING 4.1 Introduction 4-1 4.2 Projected Corrective Maintenance and 4-1 and Expected Component Life 4.2.1 Cylinder Heads 4-1 4.2.2 Cylinder Liners 4-3 4.2.3 Pistons 4-4 4.2.4 Master Link an'd Connecting Rods 4-4 4.2.5 Camshafts 4-6 4.2.6 Main Bearings 4-6 4.2.7 Cylinder Block 4-6 1 4.2.8 Major Overhauls 4-8 4.2.9 Turbochargers 4-8 1 4.3 Additional Modifications and Corrections 4-9 of Problems Created by Engine De-Rating 4.3.1 Lube Oil Systems 4-9 i
TABLE OF CONTENTS CONTINUED Section Page 4.3.2 Cooling Water System 4-10 ;
4.3.3 Turbochargers 4-10 4.3.4 Waste Heat Boiler 4-10 4.3.5 Engine Performance Optimization 4-11 5.0 RE-ENGINING ECONOMIC ANALYSIS AND COMPARISON 5-1 0F HISTORICAL MAIN ENGINE OPERATING COSTS AND EXPECTED DE-RATED ENGINE OPERATING COSTS 5.1 Historical Main Engine Related Cost 5-1 Review and Development 5.2 Propulsion System Modifications Required 5-4 In Addition To Or As A Result of Main.
Engine De-Rating 5.3 Re-Engining Economic Trade-Off Analysis 5-5 5.3.1 Cost Elements 5-5 5.3.2 Economic Analysis Methodology 5- 1-2 5.3.3 Sensitivity Analysis 5-17 5.4 Discussion of Results 5-17
6.0 CONCLUSION
S AND RECOMMENDATIONS 6-1 APPENDICES APPENDIX A: Trial Agenda, M.V. COLUMBIA, March 24, 1983 APPENDIX B: M.V. COLUMBIA, March 24-25, Trial Data APPENDIX C: M.V. C01PMBIA Shaft Horsepower Measurement, Sea Triats, March 24-25, 1983 APPENDIX D: BMEP Formulae and Sample Calculations APPENDIX E: Fuel Rate Calculations and Fuel Analysis Report APPENDIX F: Turbo and Engine Air Exchange Data and Sample Calculations APPENDIX G: Turbocharger Combined Efficiency Formulae and Sample Calculations APPENDIX H: M.V. COLUMBIA Main Engine Corrective Maintenance Tables in Chronological Sequence APPENDIX I: Economic Analysis Computations ii A
LIST OF FIGURES Figure No. Page 2.1 Fuel Rack vs Engine Speed DMRV-16-4 2-6 With C-17 Turbos 2.2 Plot of Trial vs Predicted SHP & SRPM 2-10 2.3 Comparison of Predicted &' Trial Speed 2-11 Power Data 2.4 Locked Rack Test DMRV-16-4-72033 2-15 9,200 HP @ 450 RPM 2.5 Comparison of M.V. COLUMBIA Starboard 2-18 Engine BSFC's, March 25, 1983 2.6 DE-C-17-123 Turbo Performance, M.V. COLUMBIA 2-22 March 25, 1983 2.7 Predicted and Observed Air Flow and Manifold 2-26 Pressure vs ERPM 4
2.8 Propeller Law Power Curves With a Controllable 2-30 Pitch Propeller 2.9 Compressor Matching With a Controllable 2-30 Pitch Propeller 3.1 DMRV-16-4 High Corrective Maintenance Areas 3-6 3.2 DMRV-16-4 Cylinder Head & Valves 3-9 3.3 Permanent Liner Deformation, Bore Diameter 3-12 3.4 Bore Diameter, Engine Blocks M.V. COLUMBIA 3-13 3.5 Master Rod & Connecting Rod Box Assembly 3-17 3.6 Upper Cylinder Liner & Block Section 3-22 3.7 Cylinder Configuration, Engine Block 3-23 3.8 Nondestructive Testing, Cylinder Block, 3-24 Shear Cracks, Counterbore Lip 3.9 Nondestructive Testing, Cylinder Block, 3-25 Delamination Cracks 3.10 Nondestructive Testing, Port Main Engine 3-26
'3.11 Nondestructive Testing, Starboard Main Engine 3-27 3.12 M.V. COLUMBIA, Operational & Maintenance 3-31 Periods 5.1 Impact of Varying Fuel Quality on Engine 5-11 Maintenance, Total Spares, Consumables and Labor i
iii
LIST OF TABLES Table No. Page 2.1 Scheduled vs Actual Engine Performance 2-4 Load Point Test Duration 2.2 Comparison of Starboard Engine Power Out; 2-7 Predicted by Fuel Rack and ERPM vs Observed Fuel Rack and ERPM and As Measured At the Shaft _By Torsionmeter 2.3 Computed BMEP's From Trial Results For 2-13 De-Rated Starboard Engine 2.4 Comparison of Various DMRV-16-4 Engine BMEP's 2-14 at Similar Loads 2.5 Comparison of Smoke Test Results, July 1981 2-17 vs March 25, 1983 2.6 Turbo Air Flow Calculations Results 2-21 2.7 Comparison of Computed and TDI Predicted 2- 14 Combined Turbo / Compressor Efficiency -
3.1A Summary of M.V. COLUMBIA Enterprise.DMRV-16-4 3-3 Maintenance / Failure History 3.1B Summary M.V. COLUMBIA Documented Component 3-4 Failure Modes 5.1 Summary of Estimated Annual Main Engine .5-3 Related M&R Cost, 1976 to 1982 5.2 Additional Propulsion System Modifications 5-6 Required for M.V. COLUMBIA After De-Rating As Of April 1, 1983 5.3 Fuel and Lube Oil Unit Costs 5-13 5.4 Summary of Acquisiti.on and First Year Annual 5-13 Operating Cost Estimates 5.5 Definition of Economic Analysis Terminology 5-14 5.6 Economic Analysis In-Put Data and Assumptions 5-16 5.7 Summary of Re-Engining Economic Analysis 5-18 Results for New Engine Operation on MD0 5.8 Summary of Re-Engining Economic Analysis 5-19 Results.for New Engine Operation on HF0 l
l.
iv
-. ~
i l
1
1.0 INTRODUCTION
AND EXECUTIVE
SUMMARY
1.1 oackground
-I n ' support of a major engineering change for the . power plant
, in the. passenger and vehicle ferry, M.V. COLUMBIA, consisting of the installation of new turbochargers and 'the de-rating of the main propulsion engines, Seaworthy Engine Systems, Inc.,
was retained by the State of Alaska, Department of Transportation and Public Facilities, Division of Marine Highway Systems, and tasked with a review of the adequacy of the de-rating and evaluation of the post de-rating trial performance of the vessel's main propulsion engines. As an additional (and related) t a s !:,
Seaworthy was also requested to review historical main en'ine g
component failures, and where available, associated costs to provide further insight as to the ultimate adequacy of the engine de-rating in terms of anticipated improvements in reliabil-ity, performance and associated operating economics.
The M.V. COLUMBIA was delivered as a combination vehicle and passenger ferry by Lockheed Shipbuilding Company in 1974 for the Southeastern Alaska / Seattle, Washington service. She is 418 feet long, overall, having an 85.13 foot beam and a depth of 24 feet.. At a full load displacement of 7745 Long Tons, 4
the vessel has a draft of 17.6 feet. The ship is propelled by a twin shaft medium speed diesel engine propulsion plant supported by three (3) 900 KW auxiliary diesel generators, a combination waste heat recovered / oil-fired steam generating system and two (2) saltwater distillers. Each main propulsion 1-1
. , -- .- - = ~ - .. _ . _ ._ - . . - - -
?
shaft is ' fitted with an Allis-Chalmers/Escher-Wyss controllable and reversible pitch propeller capable of delivering a maximum of 9900 HP at 250 SRPM, driven by a single engine through a 1.8:1 ratio single stage reduction gear.
The two (2) V-type turbocharged main engines are DeLaval-Enterprise model DMRV-16-4 units (serial nos. 72034 Port, 7 203'3 S t bd . ) ,
each capable of developing a maximum of 9200 BHP at 450 RPM (prior to de-rating).
1.2 Executive Summary Scope / Objective: To evaluate the historical operation and mainte-nance and repair of, and the recent de-rating of, the main
~
engine in the State of Alaska Vessel, M.V. COLUMBIA, by the completion of the following tasks:
- 1. Observation and evaluation of the ve s sel's sea trial after de-rating, held on March 24-25, 1983.
- 2. Review and summarize historical main engine component failures to date and related maintenance and repair records, including cost data, where available.
- 3. Analyze and review the existing engine de-rating to identify and quantify, where possible:
l i
- Adequacy of the de-rating modifications
- Additional modifications required to ensure engine reliability
- Cost in time and dollars to make additional modifications
- Engine life expectancy once de-rating and additional 1-2
required modifications are completed
- Cost effectiveness of re-engining the M.V. COLUMBIA
, versus continued operation of the de-rated engines.
Supporting Documentation /Results: The method of approach, support-ing documentation and data and results of the completion of the required scope of work are presented in detail in Sections
+
2.0, 3.0, 4.0 and 5.0 and the Appendices of this report.
Conclusions and Recommendations: Supported and substantiated by data and documentation contained in preceeding sections of the report, the following pertinent conclusions and recommenda-tions have been extracted from Section 6.0.
- Sea Trial Performance:
- 1. The engines as de-rated by TDI failed to develop the required power outputs as specified in the work scop'a of the contract authorizing this work.
- 2. The turbochargers, as indicated by surge problems observed during the trials and on subsequent voyages, are not properly matched to the new de-rated engine operating profile. Emperical data presented in Section 2.0 further supports this conclusion.
- 3. Numerous other problems of a smaller magnitude also identified in Section 2.0, have developed as a result of the de-rating vork and for the most part are unresolved.
1
- 4. Adequate air flow appears to have been provided tc the engines by the new turbochargers. Brake Mean Effective i
1-3
Pressures at the new operating outputs 'are equal to, or less than, those specified in the de-rating contract.
- 5. It is mov'.alc that some minor portion of the turbo-charger surg. problem is related to the difficulties being encountered with the pitch scheduling portion of the main engine control system. TDI should be required to assist and work closely with Mathers Controls to establish responsibility for and correct this situation.
- 6. Based on the above described performance, TDI should be put on notice that the de-rating work to date ,
is unacceptable and payment withheld.
- Adequacy of the Engine De-Rating:
- 1. Based on a review of main engine historical maintenance and repair data and a comparison of engine component failure frequency and mode with the modification accom-plished as a result of the de-rating effort, it is anticipated that only minimal overall improvement in failure rates and time between failures or overhauls will occur. The most significant portion of this improve-ment will occur for those components directly impacted by the improved combustion process which results from the increased availability of air blown for combustion.
- 2. It is believed that for the remainder of the engine component failures identified in Sections 3.0 and 4.0, those not dire'ctly influenced by increased air flow, little or no change in failure rate, and probably no 1-4
more than would be obtainsd by simply running the original engines at a redaced output without officially de-rating, will occur. These component failures include:
- Cylinder heads - design and manufacturing defects
- Cylinder liner distortion and wear - due to block distortion
- Piston ring distortion and wear - due to block distortion
- Cylinder blocks - distortion'and cracking
- Connecting rod bearings - design of articulated connect-ing rod assembly
- Main bearings - premature wear, high loading
- Camshafts - premature wear
- 3. It is estimated that when equated to dollars, the reduction in main engine maintenance and' repair histor-ical average annual cost resulting from de-rating may approach twenty-five percent (25%).
- 4. The existing de-rated engines,' after incorporation of the additional modifications identified in this report, can be kept running almost indefinitely if AMSH is willing to continue to maintain them at the same expensive rate, in terms of time and money.
- Additional Modifications:
- 1. Numeroun additional modifications have been identified in Section 5.0 and should be incorporated to enhance the future reliable and efficient operation of the 1-5
(
ds-rated . entinas. Somm of tha more important of thase modifications are a result of, and not in addition to, the de-rating effort. The most significant of these is the turbocharger mismatch which should be rectified by TDI by installing new matched turbochargers at no. additional cost to the de-rating contract...
- Economic Evaluation of Re-engining of the M/V COLUMBIA:
- 1. Re-engining of the COLUMBIA for operation on Marine Diesel Oil, MDO, depending on the acquisition cost estimate / remaining vessel life combination conside red ,
can offer a significant-economic advantage over continued operation of the existing de-rated engines on MDO.
- 2. Re-engining of the vessel to operate on Heavy Fuel Oil, HFO, is a clearly superior economic alternative compared to both re-engining for MDO operation or continued operation of the de-rated engines on MDO, regardless of the acquisition c'ost/ investment period combination considered in the economic analysis presented in Section 5.0.
Based on the technical aralysis and evaluation conducted and documented in this report and the results derived for the range
! of estimated re-engining acquisition cost /remcining vessel life combinations considered as part of the economic analysis, l
it is recommended that the M/V COLUMBIA be re-engined for HFO operation at the earliest opportunity.
L-6 r
L
- 2.0 ANALYSIS OF DE-RATED ENGINES AND NEW TURBOCHARGERS i
! The assessment of engine and turbocharger performance after de-rating and installatilon of the new turbos is based on design and shop test data provided by Transamerica DeLaval, Inc. (TDI)
,_ and sea trial observations and data obtained jointly by TDI, .
Seaworthy and AMHS personnel during unde,rway tests on the COLUMBIA i on March 24 and 25, 1983. Subsequent operating problems reported r i by the crew (up to the time of report preparation) during the I
vessel's initial voyage of the season, commencing April 1, 1983, are also commented on in this section. Briefly, the scope l
of work to be carried out by TDI as a part of the engine de-j rating process or in conjunction with this work and which impacts t
l engine / turbo performance, as defined in State of Alaska Delivery a
j Order 707573 (Reference No. 1) included:
l 1. De-rating of the main propulsion-engines from 9200 BHP /
l 450 ERPM each to the following operating conditions and j limits: .
i j - Idle Speed: 300 ERPM
- Design Service Rating: 5284 BHP @ 384 ERPM l - Maximum Continuous Rating: 6164 BHP @ 403 ERPM
) - 107. Overload Rating: 6791 BHP @ 403 ERPM i
i t 2. Reduction in brake mean effective pressure from 21 3 PSI
] to approximately 158 PSI.
t
! t j 3. Procurement and installation of new DeLaval-Enterprise
? C-17-123 turbochargers, two (2) per engine, four (4) total, i
l 1
i l '
l 2-1 l
having th2 following parformance characteristics:
Response time from 40 to 100% load; 6 to 7 seconds Response during rapid propeller de-pitching (engine unloading) to be at least as good as the original Elliot G-90 series units being replaced.
- 4. Installation of a Trabon lubricating oil system to -seal against exhaust valve stem and guide soot and exhaust gas blow by.
The anticipated improvement in engine performance and reliability resulting from the incorporation of the above described modiffca-tions would reasonably be expected to be manifested by reducticns in:
smoke level combustion chamber deposits lube oil contamination cylinder liner wear exhaust valve / guide blow,by The discussion of various aspects of the de-rated engine and new turbocharger performance in the following paragraphs deals largely with the results of various computations and comparison of data obtained from the previously mentioned sea trials and design or ship test data provided by TDI for the installed and/or comparable engines and turbos. While these results are felt to be directionally indicative of current engine and turbo performance, the absolute values shown in certain instances j should be viewed with some reservation due to the nature of l
! the trial data obtained and the available engine design and l
2-2 u_
oporating -basalina comparetivo performanca information. These qualifications are summarized briefly below.
- March 24-25, 1983 Sea Trials:
Due to the lateness in completing the work associated with engine de-rating and the limited time available to plan and establish rigorous trials, test- procedures and install test equipment, the testing performed was quite cursory and unusually brief. (See Appendix A for Trials Agenda).
Only the starboard engine ' and its shaft line were inspru-mented. As a result, data and calculations have reasonably been assumed as typical for both engines.
Actual sea trials were compressed from a time standpoint.
Thus, various tests were conducted simultaneously with 1 '
or at the expense of others. For example, pitch / load control systems test and adjustments were conducted i
simultaneously with steady state ' power runs for engine performance evaluation. Difficulty with the control system actually caused certain runs to be aborted or shortened. In general, the time alloted for data gathering at each engine load point was felt to be les.s than desirable (see Table 2.1).
, - The fuel oil meters fitted on the starboard engine for the test were of questionable accuracy, despite an attempt t$y Todd Shipyards to calibrate them prior i
to trials o.n March 25, 1983. A better selection could l have been made if adequac.e time had been available.
2-3
TABLE 2.1 SCHEDULED vs ACTUAL ENGINE PERFORMANCE LOAD POINT TEST DURATION-LOAD POINT SRPM/ERPM/ BHP SCHEDULED TIME ACTUAL TIME (COMMENTS) 167/300/2500 1 Hour 3/24/83: 1 hour1.157407e-5 days <br />2.777778e-4 hours <br />1.653439e-6 weeks <br />3.805e-7 months <br />, maneuvering, engine break-in 184/330/3300 1 Hour 3/25/83:1 hour, 5 minutes.
F.O. meters out of calibration vessel turning frequently. Test-ing halted due to control prob-lems and port engine intercooler 3
transition ducting leak.
202/363/4300 1 Hour l 3/25/83: 1 hour1.157407e-5 days <br />2.777778e-4 hours <br />1.653439e-6 weeks <br />3.805e-7 months <br />, Seawater cool-ing system on hand contro'l (off/
on) due to problems with thermo- i static control valves in various cooling loops which continued !
throughout the trial,F.O. meters recalibrated on evening of 3/24/
83 by Todd.
214/385/5300 i Hour 3/25/83: Test aborted and re-started twice due to maneuvering ; ,
requirements and engine control -
system problems over a two hour period. Only final 25 minutes are felt to,be representative of steady state operation.
224/403/6200 4 Hours 3/25/83: 1 hour1.157407e-5 days <br />2.777778e-4 hours <br />1.653439e-6 weeks <br />3.805e-7 months <br />, 40 minutes.
(Maximum Continuous Cut short upon return to dock.
Rating) 224/403/6800 1 Hour Not Run. Engine could not reach (10% Overload) overlaod at 403 ERPM with pro-peller on maximum pitch.
- Design Predicted Performance Data:
Apparently there is no TDI published standardized fuel l consumption map of brake specific fuel consumption rate versus BMEP or BHP and speed. (P.equested by Seaworthy).
l 2-4 y ,- --- -, -----,w -- ,y,. -
-n-,,
C-17-123 turbochargor cnd cngina p2rform2nce for the COLUMBIA's de-rated engines is predicted with no shop test comparative basis available.
Rack setting versus engine speed and power was predicted and, by necessity, not confirmed on a test stand prior to trials.
There appeared to be little coordination between TDI and Mathers Controls prior to trials relative to integra-tion of pitch control schedule with the performance characteristics of the new turbos and de-rated engines.
2.1 Engine Performance Engine performance evaluation as discussed here consisted of a review and comparison of data obtained during the sea trial with predicted values or test stand information from TDI and past performance information for the original engine configuration at similar outputs. Specifically, engine power output correlated satisfactorily with predicted and observed vessel speed / power data, brake mean effective pressure, fuel consumption and apparent combustion quality are addressed in the following paragraphs.
Sea trial data gathered on March 24-25, 1983 by both TDI and Seaworthy is contained in Appendix B.
Power Output: Figure 2.1 presents plots ,
of brake horse awer (BHP) produced by the de-rated main engines, including power as predicted by TDI versus fuel rack setting and ERPM, power determined as a result of average rack setting observed during trials versus ERPM and a final curve of BHP developed, corrected 1
2-5 '
I 1
Figuro 2.1 Fuel Rack vs Engine Speed DMRV-16-4 with C-17 Turbo M.V. COLUMBI A id o m I t 8 t t- - '
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from trial shaft horsepowar (SHP) measurements m:da by Seaworthy assuming a 987.- gearing and shafting mechanical efficiency.
(The theory of operation of the torsionmeter utilized and sample traces of recorded data are contained in Appendix C). Table 2.2 presents a comparison of these values at test points from 360 ERPM and above. The original predicted load curve of fuel rack versus BHP and ERPM assumes the propeller law's approximate full pitch cubic relationship. During the trials, very close to full pitch was applied to the propellers from 360 ERPM on up, equating to somewhere between 1.0 and 1.1 pitch to diameter l
l (P/D) ratio as reported by Mathers Controls personnel who cere l onboard testing and adjusting the engine controls during this period.
i TABLE 2.2
- COMPARISON OF STARBOARD ENGINE POWER OUT; PREDICTED BY FUEL RACK AND ERPM VERSUS OBSERVED FUEL RACK AND ERPM AND AS MEASURED AT THE SHAFT BY TORSIONMETER, MARCH 25, 1983 PREDICTED BY OBSERVED BY TORSION-ERPM/SRPM(II FUELRACKg{ I AVERAGE FUEL RACK METER (3) 363/202 25.55/4400 BHP 24.1/3750 BHP 3640 BHP ,
385/215 26.7/5250 BHP 24.9/4250 BHP 4680 BHP I
403/224 28.3 /16200 BHP 25.85/4900 BHP 5930 BHP (1) Propeller pitch at maximum, 1.0 i P/D 6 1.1 (2) From Figure 2.1 (3) Corrected from measured SHP values assuming a 98%
l gear / shafting mechanical efficiency i
I l
2-7
From cn insp2ction of th2 dete presented in Figure 2.1 end Table 2.2, neither pot.2r determined from actual fuel rack setting and ERPM or n m.ssure'd at the shaft very closely matched the predicted engine load profile, shown in Figure 2.1. Further, at all test ERPM's but 363, TDI observed rack BHP falls below that determined from measured SHP. Relative to contractual performance, TDI test data and resultant plotted Brake Horsepower data fails to meet anticipated outputs for Design, Maximum i Continuous and 10*/. overload service ratings of 385 ERPM/5284 BHP, 403 ERPM/6164 BHP and 403 ERPM/6791 BHP, respectively.
Additionally, oer the torsionmeter, the Maximum Continuous Rating of 6164 BHP at 403 ERPM appears not to have been met based on Seaworthy's measured SHP data at this load. Also, the 10% overload capability at 403 ERPM could not be demonstrated as the propellers were on full pitch from at least 360 ERPM on up. Thus, the only way that load could have been increased was to increase engine speed above the dew limit of 403 ERPM established for the de-rated engines. Further, with the propellers on full pitch, engine / shaft RPM and power output would be expected to more closely follow the propeller law cubic speed / power relationship as is the case wi.th the speed / power curve plotted from measured SHP and ERPM, as compared with the speed / power l curve obtained utilizing average . fuel pump rack setting and observed ERPM. Also, the f'uel rack / BHP /ERPM load curve shown by TDI in Figure 2.1 is a predicted one, never previously verified by actual tests for the COLUMBIA's engines. In addition, it i
I l
2-8
is noted here thnt the adjustmant and celibration of the starboard engine's number. one right bank fuel pump and rack assembly had been altered and never reset prior to the trials. This fuel rack position is used as the master command signal indicator for that engine's speed and load control program. In consideration of the previously discussed factors, the data obtained from the torsionmeter reading is felt to be more closely representative of actual power produced.
Vessel Speed: To further evaluate and verify the speed and power relationships derived from the test data from the shaft torsionmeter and from fuel rack settings, a comparison of vessel speed over the ground taken for each test run from the bridge was plotted versus rack and torsionmeter power outputs and compared with predicted vessel speed / power curves for propeller P/D ratios of 1.0 and 1.1, as presented on pages 19 and 20 in Morris Guralnick Associates, Inc. report, " Performance Predictions and Engine Selection Criteria for the M.V. COLUMBIA",
dated June 1982 (Reference No. 2). The results of these comparisons are shown graphically in Figure 2.2 and 2.3. By inspection of Figure 2.2, the shaft horsepowers and SRPM's plotted for the torsionmeter data are much more consistent with the shape of, and. fall very closely to, the predicted P/D = 1.0 curve while the plot of the rack determined SHP versus SRPM falls i
well below the P/D = 1.0 line, which would indicate a P/D tatio of less than 1.0 in contradiction to the pitch carried during the trials as reported by Mathers. Trial and predicted data W
2-9
Figuro 2.2 Plot of Trial vc Prodictod SHP cnd SRPM, M.V. COLUMBIA F- -
22 N __ 2
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Figuro 2.3 Comparison of Predicted and Trial Speed Power Data, M.V. COLUMBIA 4 - - - - ~ -
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(Ship Speed in Knots) 2-11 _ _ _ __ . _. _-
platecd for vascol spacd and SHP for P/D ratics is shown in Figure 2.3. Here also, the observed uncorrected speed /SHP points recorded during the test from the to'rsionmeter show a much closer agreement with vessel speeds predicted for P/D ratios of 1.0 'and 1.1 than speed /SHP data based on and plotted for fuel rack settings.
The predicted speed, SRPM and SHP data plotted in Figures 2.2 and 2.3 was extrapolated (in Reference No. 2) from model test data and initial delivery sea trial data. It also contains adjustments for estinates of increased hull roughness as a r
function of time out of drydock and additional wetted surface areas which would result from a planned lengthening o f- the existing skeg to improve the vessel's manuevering characteristics .
It is estimated that these adjustments increased required SHP by three (3) to four (4) percent over what would be the case for the hull at the time of testing on March 25, 1983. This results' from the fact that the fact that the hull was freshly painted and that the skeg had not been lengthened. The ship speed and shaft horsepower data recorded during the test was taken with the vessel operating at a draft of 13'5" FWD and 16'0" Aft, resulting in a mean draft of 14'8.5". Thy draft on which predicted speed and power curves were based was 16.0 feet even keel. Again, based on a comparison of the observed trial vessel speed / power data with predicted vesse? speed / power i data, the BHP's determined from the torsionmeter readings appear to be more closely representative of actual power produced l than do the equivalent rack setting values of BHP.
i 2-12
Broko Macn Effective Pressures: BrckG M2mn Effective Pressures lL L
[\ .were calculated for BHP's as determined from fuel rack and i ERPM data and from SHP and ERPM data recorded at 363, 385 and i
! 403 ERPM on March 25, 1983. Formulae and sample calculations j .are presented in Appendix D. The results of these calculations are contained in Table 2.3.
I TABLE 2.3 COMPUTED BMEP'S FROM TRIAL RESULTS FOR -
DE-RATED STARBOARD ENGINE, MARCH 25, 1983 I
PREDICTED RACK III OBSERVED RACK (1) TORSIONMETER(2)
ERPM BHP /BMEP, PSI BHP /BMEP, PSI BHP /BMEP,' PSI l 363 4400/125.8 3750/107.2 3640/104.1 385 5250/141.5 4250/114.6 4680/126.2 I 403 6200/159.6 4900/126.2 5930/152.8 A i
i (1) From Figure 2.1
- l (2) 98% gearing and shafting mechanical efficiency assumed i
1
{ As can be seen from the results, computt,d BMEP's, regardless 4
of power measurement results utilized, did not exceed the maximum ,
j 4
limit set by the de-rating contract workscope of 158 PSI.
p Further, a review of COLUMBIA's starboard engine test stand data and June 1981 sea trial data contained in TDI's. report, 4
" Shipboard Test, M.V. COLUMBIA, Starboard Engine, S/N 72033",
l 1 August 31, 1981 (Reference No. 3) and test stand data for a i
similar DMRV-16-4 engine (Reference No. 4) was conducted to
- determine BMEP's at ERPM's similar to those run during the j March 25, 1983 test. A review of past engine room log data .
j l 1
1 2-13
fcr th2 COLUMBIA's secrboard ongina w33 clso parform:d in en attempt to establish a typical load profile from which BMEP's could also be computed. The results of these' investigations are presented in Table 2.4.
TABLE 2.4 COMPARISON OF VARIOUS DMRV-16-4 ENGINE BMEP'S AT SIMILAR LOADS
-. MARCH 25, 1983 TRIAL ERPM/ BHP /BMEP ENGINE 363/3640/111.5 385/4680/135.2 403/5930/152.8
- 1. Halter Maring Test 390/2366/63 390/5521/%47 390/6704/178.5 Stand, 12/78tl) I
- 2. COLUMBIA: Sebd 320/4089/137.2 360/5814/167.7 400/6460/167./
Engine Test Stand 7/72(2)
- 3. COLUMBIA: Stbd 347/2700/80.8 368/3950/127.3 401/7270/194 Engigg)SeaTrial,
- 4. COLUMBIA: Stbd 399/7500/204,G Engine (2) 5-6/81 ----- -----
399/7500/194 .
7/80 ----- -----
396/7500/196.6 7/77 ----- -----
400/7300/189.4 6/74 ----- -----
430/7400/178.7 (1) With DE C-17-123 turbos (2) With Elliot 6-90 series turbos BMEP's for various COLUMBIA voyages were computed from log book ERPM and rack settings per the July 1981 sea trial load curve for the original engine rating shown in Figure 2.4 and taken from Reference No. 3. A comparison of BMEP's shown indicates the following. First, it appears that the engine has been operated full away at combinations of ERPM, propeller pitch and BHP l
l which result in BMEP's ranging from approximately 175 to 200 2-14
~
Figuro 2.4 Locked Rack Test DMRV-16-4~-7 2033 .
9,200 HP@ 450 RPM (Prior To De-rating) 00/) . . I ! *. .
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PSI as ccn , bo best determined utilizing Figure 2.4 (However, it should be noted that frequent changes in the pitch program portion of the throttle control system, as reported, make it difficult to conclude that the load curve shown in Figure 2.4 is 100% representative of engine load profile from delivery in 1974 up until de-rating in early 1983.) Secondly, from Tables 2.3 and 2.4, operation at ERPM's and BHP's, as shown from the COLUMBIA and Halter Marine engine test stand data, which are somewhat similar to the predicted de-rated engine load profile, should produce BMEP's which are ten (10) to twenty (20) percent lower at the projected maximum continuous rating of 6164 GHP for the de-rated engines. Operation at the new Design Service Rating of 5284 BHP /355 ERPM should result in a reduction of 8
from twenty (20) to thirty (30) percent in BMEP's compared to past operating loads.
Combustion Quality: As a qualitative assessment based on smoke and particulate emission determined from a Bosch smcke test apparatus and visual observation of the stack at various visady state loads, it appears that the combination of new turbochargers and the engine de-rating have significantly improved the combustion process. Stack emissions were virtually clear up to the maximum load point at 403 ERPM where a very sligh't haze was observed.
Further, data shown in Table 2.5, which compares the smoke results from the July 1981 sea trial with those taken on March l
25, 1983, also indicates a substantial reduction in visable l
l smoke.
i i 2-16 m,. -
,_ w a
i TABLE 2.5 COMPARISON OF SMOKE TEST RESULTS '
JULY 1981.vs MARCH 25, 1983 JULY 1981 MARCH 25, 1983 332 ERPM/0.5 BSN 363 ERPM/0.3 BSN 367 ERPM/0.4 BSN 385 ERPM/0.225 BSN 401 ERPM/0.8 BSN 403 ERPM/0.3T BSN The reduction in exhaust gas smoke level, while indicative of an improvement in combustion quality, cannot be utilized as an absolute indicator of combustion efficiency or the comple.te-ness with which the potential chemical energy in the fuel is converted to heat via combustion in the engine's cylinders.
It is possible to h' ave a significant amount of fuel in various stages of oxidatien exit in the cylinders with the exhaust gases in a clear state, if sufficient air is being supplied by the turbos.
Fuel Consumption: Figure 2.5 presents plots of brake specific fuel consumption rate (BSFC) in LBS/ BHP-HR for various conditions.
Briefly, BSFC can be viewed as an indicator of how efficiently an engine converts the energy in a pound of fuel to a unit of power, the lower the BSFC the more ' efficient the engine.
First, as. shown, test stand BSFC curves for the COLUMBIA's i i
starboard engine and a similar newer DMRV-16-4 engine delivered to Halter Marine are plotted from data contained in Reference No. 3 and No. 4 and show excellent agreement with fuel rates j as predicted by the Builder. Utilizing . fuel flow and various 2-17
Figure 2.5 Comparison .of M.V. COLUMBI A STB.D Engine BSFCs, 3/25/83, with Original Test Stand Data
.55--
M.V. COLUM8I A-1/23
.50- -
- J J /83' MEIELTEST OKTA
..'8fif BY.
- ~
RACK._ FfG'. 2.1 x ,e ~ ~a G<\ .
S.V. COLUMBIA
/p 3/25/83f.'07. YEST40ATA, METER 7.IOR$10HsETER
. 4 5- - -- - - . ;. . . . -
cc Y
a.
'x \ -
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% \ X co .4 0 -
cc J '
O HALTER M ARINE-0 MRV-16-4 TEST ST ANO DATA,
.35- - -
- " -~
-a- -
, n R.V. COLUMBIA TEST STANO DATA.
l 7/72 (REF; NO. 3) ST80 ENGINE M.V. COLUMBIA
- 30 . . . . . _ ... . _. . . _ . _ . . . _ . . _ _ _ . . ._ 3/23/83 TEST'OATA, 8HP-8T RACK- FIG. 2.1 i
, FUEL BY. RACK.CREf...NO. 3)
! .25 ; ! ! l [
2000 3000 4000 Sooo sooo 7000 l
, BRAKE HORSEPOWER l
2-18 1 3-.hdM, a..q.g.W #ma"
, y g , --y, . ,- - - -
w - ,, -- - -
l power' data sources obtained during the sea trials, curves repre-sentative of actual in-service performance have also been plotted 1
in Figure 2.5. Two curves based on fuel consumption from the i test meters and TDI fuel consumption as predicted by rack setting (Reference No. 3) and TDI test de-rated engine brake horse-power from rack and ERPM (Figure 2.1) were also plotted. These curves are the upper and lower most lines on Figure ~2. 5. They show virtually no agreement between either the predicted fuel flow or the flow as measured, one being 45 percent higher and the other 14.7 percent lower than original test stand BSFC values shown in Figure 2.5. The fifth and final curve pl,ots BSFC for the vessel's starboard engine from data obtained from the test fuel oil meters and from power as measured by the torsionmeter installed for the trial. This curve shows a much greater ' slope than the test stand data, with fuel rate decreasing with increasing engine load. From this curve, at 5815 BHP, the difference between the test stand performance and the observed BSFC for the starboard engine is 18%. 'A plot of BSFC based-on power from the torsionmeter and fuel from the rack setting
(
would, in fact, result in a curve that would fall well below the abscisa of Figure 2.5. As in the case of the TDI data plot, resultant fuel rates in th,e r.ange of- 0.233 to 0.30 LBS/ BHP-HR, equating from 45% to 62% thermal efficiency, are well outside the range of the most efficient medium speed diesel engine capability and therefore, are unacceptable. BSFC data derived utilizing fuel meter flow rates in both instances (TDI predicted 2-19 i
. -_ , . . _ _ , _ . ~ -_- -. _
l and torsionmotor powar outputs) show brake specific fuel races ;
considerably in excess of the factory test stand rates. As stated in the introduction to Section 2.0, the absolute magnitude of fuel flow values recorded by the test meters may be open to challenge. Based on the past history of these engines, it would seem reasonable to assume that they are in fact consuming fuel at a rate considerably in excess of original and design predicted performance, perhaps by as much as 107.. Potential sources of this increase may include operation at reduced ERPM, cylinder load imbalance, improper fuel injection timing, lack of an optimized fuel metering system (nozzle, injector, pump) for low load operation, increased cylinder liner / piston clearances, reduced BMEP and less than anticipated turbocharger efficiencies.
All fuel rates shown have been corrected to design on the basis of lower heating value content of the f6el actually burned to the design lower fuel heating value content of 18,190 BTU /LB assumed for design predicted performance , calculations. Sample fuel rate calculations are contained in Appendix E, along with a laboratory analysis of the fuel actually burned during the trial.
2.2 Turbocharger Performance Turbocharger performance was reviewed quantitively and qualitative-i ly based on data obtained for the starboard engine from the i
" arch 25, 1983 trials. Comparisons have also bein made with the original Elliot turbochargers, based on data contained in Reference No. 3.
l 2-20
Air Delivered: Table 2.6 presents the results of turbo and 1
engine air an'd gas computations which quantify the observed
, air flows delivered by the new DE C-17-123 turbos. Data, formulae
! p &
and sample calculations are contained in Appendices B and 5 TABLE 2.6 TURBO AIR FLOW CALCULATIONS RESULTS AVG. TURBO ENGINE AVG. COMPRESSOR AVG. TOTAL AVG. TOTAL RPM A/F RATIO PRESSURE RATIO ACFM. PER SCFM PER ERPM TURBO TURBO 363 12,240 28.67 1.725 7100.8 67J1.7 385 13,608 30.54 2.03 8956.5 8483.3 403 15,100 30.71 2.245 10273.4 9747.3 f
The results of thir tabulation have also been presented graphically in Figure 2.6 in which corrected air flows in SFCM have been plotted versus compressor (boost) pressure ratio for the average values shown in Table 2.6 and for the, individual right and left bank blower outputs. The correction from actual to standard flow (SCFM) was made to take into account the compressor inlet temperature and pressure difference between the conditions observed on the vessel and the design standards on which the 4
unit's design perfo'rmance is based, as shown in Figure 2.6.
The data, as plotted in Figure 2.6, would indicate that the compressors are operating considerably closer to the predicted ,
l surge line than would be desired as shown by the relation of ;
the lines of observed performance which fall to the outside l 2-21 l
4
i t
Figuro 2.6 DE-C-17-123 Turbo Perf ormance, 3/25/83 M.V. Columbia s... v....... ... C.,.. .. 5. n 1,... -._ .....
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'and to the left of the preferred engine operating band also I l
shown on. Figure 2.6. The. right bank blower pressure ratio / total air flow curve is closer to the predicted surge line than t'n e left bank plot. Subsequent to the sea trials and during initial j voyages, a turbo on the starboard engine was observed going into surge. It is most likely that this is the same turbo identi-fled as the right bank turbo by the sea trial data, as indicated by this unit's plotted performance falling closest to the theoreti-cal surge line in Figure 2.6. Therefore, it must be concluded from this data that the turbos as supplied by TDI are not properly matched to the engines' new de-rated output.
Another observation on sea trials relative to the turbos and verified during data reduction is the disagreement in plotted flows and pressure ratios in Figure 2.6 and apparent turbo RPM. The turbo RPM's logged in Appendix B by TDI, even after correcting for observed temperature, do not correlate at all with predicted RPM's on Figure 2.6. At tha time of trials there was some question as to the accuracy of the turbo tachometers supplied as part of the de-rating workscope. (Two (2) tachometers failed during the trials.) It would appear that the turbo tachometer readings are in error.
Combined Efficiency: In an attempt to provide an additional correlative data point for the turbo compressor plots shown in Figure 2.6, a combined turbo / compressor efficiency was computed for each ERPM test point. The results are presented below in Table 2.7. Da ta , f ormu,lae and sample calculations are contained 1
2 _ _ _ _
in Appandices B, F and G.
TABLE 2.7 COMPARISON OF COMPUTED AND TDI PREDICTED COMBINED TURBO / COMPRESSOR EFFICIENCY LE BANK TC RIGHT BANK TC AVERAGE TC i ERPM ( Cal]/
c Pred2/a ,7,3 ). (Cale/Pred/a ,7.) (Calc /Pred/o 37. )
363 60.87%/61.5%/1.02% 58.39%/60.0%/2.7% 59.63%/60.75%/1.86%
385 62.55%/63%/.55% 61.27%/61.8%/.85% 61.96%/62.4%/.7%
403 63.60%/64%/.625% 61.11%/62.2%/1.75% 62.36%/63.1%/1.19%
1 Calculated from test data 2 From Figure 2.6 3 a ,7. = (Pred-Calc) (100) tFred) l The results presented in Table 2.7 show a very good correlation between computed values of combined turbocharger efficiency and predicted efficiency based on the operating lines plotted for the right, left and average turbocharger values of compressor pressure ratio and corrected air flows i'n Figure 2.6. Due to the lack of accurate turbo RPM values , these data become signifi-cant in that they provide a well established third reference point which supports the location of the turbo operating lines, as plotted in Figure 2 .' 6 , closer than would be desired to the theoretical surge line for the DE C-17-123 compressors.
l Comparison of Other Performance Data: Other engine and turbo data was reviewed and compared with starboard engine performance with the original Elliot turbochargers at similar loads. The l results of these investigations which also indicate that a significant increase in airflow has occurred, are summarized below.
2-24 m n -
- Cylinder Exhaust Temperatures: A comparison of pre and
. post DE turbo installation cylinder temperatures, based on the March 25, 1983 sea trial data and data contained in Reference No. 3, indicates average temperature reductions in the rang'e of 75 to 125 F per cylinder at similar loads.
- Charge Air / Exhaust Manifold Pressure Differentials: In a gross sense, if the engine is considered as an orifice, then the pressure drop across the engine from charge air to exhaust manifold is approximately indicative of air flow through the engine. This pressure differential, after installation of the DE turbos, increased by as much' as 7.5 times at similar engine loads.
- Firing Pressures: A comparison of the March 25, 1983 sea trial data and similar information from Reference No.
4 3 shows little or no change in peak cyclinder firing pressure and continued unbalance from cylinder to cylinder. The TDI representatives onboard at tha time indicated that these were lower than anticipated and that correction of this problem by advancing the fuel injection timing and balancing the cylinder pressures in the starboard engine would likely improve overall operating efficiency.
- Charge Air Manifold Pressures: Figure 2.7 presents a plot of at.uicipated charge air manifold pressures versus ERPM l sr-provided by TDI for the new turbos. Overlaid on this graph are additional curves which plot actual manifold pressure observed versus ERPM (dash-dot line) and computed total l engine air flows in SCFM versus ERPM (dashed line). The 2-25
, 6 k
! Figure 2.7 i
t Predicted and Observed Air Flow and Manifold Pressure .
1 l
vs ERPM for DE -C-17-123 Turbos on M.V. Columbia
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predicted air flows corresponding to the observed manifold ;
1 pressures are significantly nigher than the computed values shown. Referring to Figure 2.6, this would have been the case had the . observed operating line for the turbos fallen within the preferred engine operating envelope with corres-pondingly higher turbo efficiencies. Also, in the case of both curves plotted from the March 2b, 1983 sea trial data in Figure 2.7, more total air flow at 385 and ~403 ERPM's is indicated than from the TDI plot .of predicted performance. However, both plots of the observed data
- indicate that air flow from the turbos appears to Eall off much more rapidly than predicted at lower engine loads.
This performance may account for the observed surging-during trials after rapid application and removal of propeller pitch (engine load) during response testing.
Turbo Response: On March 25, 1983, brief quetlatative tests-of turbo response to rapidly increasing and decreasing engine load commands were conducted. These consisted primarily of bridge control initiated crash astern and crash ahead maneuvers. On one such maneuver, the port engine stalled and dropped off the line completely. At various times under severe load application or - removal , all turbos were heard squealing or barking back. Some squealing, indicative of turbo surge, was also noted during steady state operation 1
(
at the 385 ERPM test point. Additionally, a very high l
l pitched noise was also determined as eminating from the i l
m 2-27
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discharge side 'of - the. turbos during the 385 ERPM test l run. It is speculated 'that this may be -the result of a harmonic,or resonant' frequency condition for the turbos ,
- -i speed, as occurring at this engine it seemed to decrease-i - .
when the engine was operated above or below this point.
It was also noted, especially by those familiar with
[ COLUMBIA's past response characteristics, that Ehe current load control program added pitch to the propellers (increased
- engine load) at a rate much higher than ever noted previously.
4 At the time it was felt that the rapid pitch application h by the control was the major causitive problem for ,the engine / turbo response difficulties previously described.
2.3 Post Trial Performance Throughout the report preparation period, and up to April 18,
- 1983, Seaworthy has been made aware of various problems and i
l conditions in .the COLUMBIA after entering service on April 1, 1983, which collaborate and expand ,
on much of the trial i
data and discussions already presented in this section.
J i
Turbo Surging: The frequency of observed turbo surging increased i
during the initial voyage, primarily on the port engine. Turbo
- surge is defined simplistically as the range of unstable operation I
which occurs when air flow through the compressor is reduced while the compressor pressure ratio (pressure at discharge divided by pressure at the _uction) remains constant, shifting the operating point on the compressor map (Figure 2.6)- to the left of the surge line. In severe cases, a flow reversal in l
! 2-28 9 - + -r 9 'ed-- y-41-y wi5-rg-y-yy+% "v' *M7'- w w P *r '" "--sr -
i the compressor may occur. Turbo surging results in unstable engine operation, air starvation, poor combustion and reduction or fluctuations in engine speed and power. Surging can also cause mechanical damage to the compressor as a result of increased mechanical stresses which occur during surge. To relieve this situation, pitch (and ERPM) were reduced. Based on the engine operating lines plotted in Figure 2.6, the problem of continued and more frequent surging is not surprising given the closeness of the engine operating line to the theoretical surge line.
Also, various changes in actual ambient conditions such as temperature, barometric pressure, humidity, intake filter clearrli-ness, etc. can cause the surge line for the turbo to shift further to the right, encroaching even more on the actual engine operating line. This is further supported by the fact that hard ship turns also cat $s ed the turbos to go into surge, giving additional credence to the closeness of the surge line to the engine operating line. However, difficulties with the load control portion of the engine control s'ystem may have also contributed to this situation.
Another contributing factor is the match of the DE-17-123 turbos capable of an output that would satisfy the air requirements of the original 9200 BHP rating of the engine. If, in fact, these are the same units in terms of capacity, they have ended up operating in a situation depicted by Figures 2.8 and 2.9, taken from Reference No. 5. Figure 2.8 shows the speed, pitch and power relationship for a generic four-cycle engine fitted Extreme with a CRP wheel. pitch seating is applied from point 2-29 i
Figure 2.8
' Propeller Law Power Curves with a Controllable Pitch Propeller 800-Monimum power curwe 600 Optimum pitch Em m /
pitch /,/ot rated speed 3 settias /.N /
400- ,
, ; /
I 5 /
Q.
/ /
200'
/
/
//
2 ,'
/ /, '
/ .- t 5 O
O 200 400 600 Speed (rev/ min)
Figure 2.9 Compressor Matching with a Controllable Pitch Propeller
- 4. 5 74 4.0-Y 70 3.5 . .80
.o 3,o- /,
L Extreme pitch
'//
52.5 - setting f '
// .80 2 2.0 - /
! t. 5 ,
f , 4s s I l.C
( 0 5 10 15 20 25 Moss flow porometer h/T/P l
l 2-30
1 (2) to (3), building up to the maximum power portion of the c u'r v e from (3) to (4) as speed is increased. If this same pitch curve is overlaid on a compressor. map for a unit matched for the engine's maximum output, as shown in Figure 2.9, the following can occur. Operation at maximum engine output, points (3) to (4), places the compressor well away from surge and close. to ,
maximum turbo efficiency. However, as load is reduced, . essentially by lowering speed, while maintaining a maximum pitch setting, points (3) to (2), the extreme pitch setting line comes very close to the surge line. This situation is further aggrevated by the " waist" or dip in the surge line characteristic of operacing a highly rated turbo at lower outputs, as shown. Thus, for an engine fitted with a CRP, it is the extreme pitch setting curve and not maximum engine / propeller speed which determines the surge margin and related matching requirements. Referring back to the data plotted in Figure 2.6, extrapolation of these operating lines to a higher compressor output shows the slope carrying them into a more stable (further frbm surge) and efficient area on the compressor map. As indicated by the dotted lines of decreasing pitch setting in Figure 2.9, a reduction in pitch setting will move the engine curve away from surge which is exactly the experience reported on the initial voyage of the COLUMBIA.
As a final point requiring clarification by TDI, relativa to turbo surge, it is noted that an increase in charge air manifold temperatures up to 50 F was desired by TDI to improve the combus-tion process at the lower operating outputs for the de-rated 2-31 1
- . - , . - . - , - - - , . ~ . . , __r_., ,,,. v - .-r y. , . --
engines. It . appears that this has been partially achieved.
, However, a review of test stand data for a similar DMRV-16--
f 4 engine-fitted with DE C-17-123 turbos (R'eference No. 4) indicated that on two occasions, at outputs of . 6027 BHP /300 ERPM and 8450 BHP /390 ERPM, ch'arge air manifold temperatures were reduced from 150 F to 125 F to eliminate turbo surge. Given the current surge problem and the test' stand data, i t would appear thac the desired increase in charge air manifold temperature for 4
improved combustion quality is a possible contributory cause of turbo surging. As a minimum, reduction in charge air temperature to reduce surging, if viable, cannot be accomplished without some negative impact on the low load combustion process.
! Trabon System: At the time of report writing it was understood that while the system was operational, certain components required t
d for proper system function, including a micro-switch, had failed.
- Proper dosage rates and frequencies had also not been provided.
Structural Items: Difficulties in this area centered around
- leaks in the compressor discharge transition piece / inter-cooler i
plenimum, specifically on the port engine outboard turbo. TDI l
had admitted that these structures have caused considerable problems as a result of cracking and leaking in similar applica-tions.
l Waste Heat Recovery System: It was reported that during the initial- voyage the oil-fired boiler operated continuously as a supplement to the waste heat boilers' steam output for auxiliary and hotel loads. This was not the case prior to de-rating.
2-32
, u. e
, , n.,. . . - - , -- y .,y, ,sw-w-e-,----- -,, p.--7.-.,.,-c- r -,-s., ---y . , - - - - ,,
- ,-vv., y- y-w.- -ve,
1 The short fall of waste heat generated steam results fron a ,
combination of factors. Operating the engines at a lower output will reduce exhaus't -gas mass flow, although this is offset somewhat by improved turbo air delivery. More significantly, the cylinder . exhaust and turbo exhaust temperatures have been substantially reduced. Thus, each pound of exhaust gas carries less heat with it up the stack to be recouped in the waste heat boiler. Because the exhaust flow after the turbo on each engine splits and flows through a silencer / spark arrester and a waste heat boiler, this situation can be remedied to some degree by diverting a greater flow of exhaust gas through 'the boiler by restricting flow through the silencer on each engine.
Cooling System: The increased jacket water temperatures desired by TDI to enhance part load or de-rated engine performance, has not been obtained. Operation in colder air and sea temperatures i in Alaskan waters on the first voyage of the season resulted in a reduction in charge air temperatures, to 145 F versus the 155 F values observed during the trial. The automatic temperature A
(AMOT) control valves were noted as functioning and closed at this time.
f Control System: Mathers Controls has been wo.rking steadily on resolving the pitch control program difficulties as reported during the sea trials and subsequent voyages. The rapid application of pitch has surely aggrevated the surging and response problems observed to date. Conversely, had TDI's predicted performance, ,
i relative to power output, fuel rack and engine RPM more closely l 2-33
cenferm;d _to what wss cctually obtain:d after de-rating and
. hrd tha-~ turbo matching been further- from the surge region,
-the control difficulties presently being experienced would probably not have been as severe or as persistent.
I e
9 2-34 i umI
3.0 HISTORICAL REVIEW AND ANALYSIS OF ENG.INE COMPONENT FAILURES 3.1 Introduction An indepth investigation of the M/V COLUMBIA's engines' operating, 1
maintenance and repair history was performed. The primary objec-tives of this investigation were:
The identification and tabulation of significant engine component failures The tabulation of major maintenance actions performed to either prevent catastrophic failures or to maintain the engine in an operational condition to meet ve rsel schedule requirements.
The identification of the causes of the failures and excessive maintenance requirements and of the resultant corrective actions taken either by the owner, if any, or TDI.
The results of this investigation are presented in the following sections with detailed supporting technical data.
3.1.1. Data Sources The following listing identifies the data sources used: I
. Chief Engineer's voyage reports
. Port Engineer's weekly reports, memos and correspondence files
. Ship's Logs American Bureau of Shipping (ABS) sarveys and files
. U.S.C.G. Reports State of Alaska, Department of Transportation files
. .Transamerica DeLaval, Inc. (TDI) reports, correspondence and invoices 3-1
Metciturgical Reports
-Lubricant' Analysis Reports Consultants' Reports and Analyses ,
The documents reviewed approximated 10,000 pages of data.
3.1.2. Time Period The engines' history war analyzed from the time of delivery, June 1974 until the 1982/83 overhaul, ending March 25, 1983.
The pre-delivery, shipyard engine tear-down was not included in the historical review.
3.1.3. Chronological Methodology The methodology used in collecting, tabulating and summarizing the data was consistent with the manner in which the Alaska Marine _ Highway System (AMHS) keeps their files. and records.
Specifically, time frames were categorized as either warm weather
" operational" periods, or " overhaul" periods when more work is typically performed. All data was evaluated on a chronological basis. The details of the chronological invest'igation are presented in Appendix B.
3.2 Maintenance History Tabulations Tables 3.lA and 3.lB provide final summaries of the data collected and analyzed. Table 3.lA provides a totalization of all significant maintenance occurrences by engine component over the operating.
Life of the engines - to date -
30,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br />. This table is set up to provide a direct comp'arison of actual component life between corrective maintenance actions and the scheduled or 3-2 him - , - - -
T A ll t. E 3.lA
SUMMARY
H/V COLUMBIA ENTERPRISE DHHV-16-4 HAINTENANCE/ FAILURE NISTORY - 30,000 ilRS/PER ENGINS
[ CORRECTIVE HAINTENANCE (( COMPONENT LIFE [
! t' $ 2 2;
.'- 2l.* 'i .i i
- 3, 2 * ? ? -
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COMPONENT 3. j j a tj = -l 4 l
- j 4 a 23 3l3:
CYLlHDER llEADS 287 2,900 8,000 12,000 20 13,000 N/A No. of Rebuilds Unknown CYLINDER LINERS 138 5,400 24,000 20,000 20 19,600 100,000 PISTONS 149 4,350 24,000 20,000 100.000 u
I w
142 5,200 24,0C0 20,000 142 5,200 ' 20,000 PISTON R I tIG C (SETS)
HASTER & LINK 97 4,390 7 16,360 '
"9 I "'
CONNECTlHG ROD e Life HASTER ROD 50' 8,035 No Listing 30,000 50 8,035 N 60,000 BEARING REPLACEHENT .
C CAH S il A FT 4 24,000 h No Listing Bearings Rolled HAIN BEARINGS ALL 16,000 24,000 20,000 ! 45,000 Top to Bottom
?
l Head, Piston, Liner Bearing, HAJOR OVERilAULS 4 6,025 24,000 20,000 l Simultaneous Removal, Re-
! .huild d egair TURBOCllARGER 40 2,340 8,000 12,000 16* 2,830 *aearings, Castings Seals TURB0CllARGER 40 2,480 8,000 12,000 20,000** " Bearings OVERilAUL
4 i
i TABLE 3 . I 11
SUMMARY
M/V COLUMBIA 30CUMENTED COMPONENT FAILURE HDDES i
6 ENTERPRISE DHRV-16 30,000 llRS/PER ENGINE l . .
- I $ =
- :
- g I IJ EI .
3 : 3 7 I: 4 s l .-u .;
COMPONENT J .! -((
- eau 6s Crac k ed 58 3,200 RebutIdable 7
CYLINDER llEADS -
M a n uf acturin g De fac ts 24 3.000 Rebul1dabIe W ar pe d He a d -
77 2,890 f ir e Ring failure Corre c ti v e M ainte n a nc e PISTONS or M o dific atio ns to 92 4.870 Croun/ Skirt I
HASTER & LINK g0NNECTINC ROD
''I!'f
'"9 40 10,080 fastener Hechaniso 49 6,475
~
expected intervals specified by the engine manufacturer, TDI, in their maintenance handbook. Additionally, data is provided for other typical medium speed engines' maintenance intervals and parts life. This data is provided to supplement TDI informa-tion in certain areas where it was lacking and to provide addition-al appropriate comparative data. The information- presented is based on typical medium speed diesels of equal or higher power and speeds, operating on MDO.
Finally, Table 3.lA identifies the number of components scrapped and the average life of that component. Again, a compartson of actual average compo'nent life to expected component life can be readily made.
Table 3.lB provides a final summary of the causes for the component corrective maintenance actions present in Table 3.lA. This table identifies the documented causes which were obtained as a result of the detailed data investigation. As can be seen, there are obvious differences in the values between the total occurrances (all causes) and the documented cause or reason totals. This is essentially due to the absence of detailed, extensive and accurate record keeping and documentation practices of the operator.
Figure 3.1 provides an illustration of the engine areas experi-encing recurring failures and inordinately high maintenance actions. These areas are identified by dark outlining.
3-5
g Figuro 3.1 U DMRV-16-4 High Corrective Maintenance Areas g 7
= 5 e/ue r _e i I <_
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i3. 3 Summary of Maintenance / Failure Isistory
' The' following section refers to Tables 3.lA and 3.lB and provides the narrative and analyzstior. of the data presented. It should be 'noted that component design is not analyzed, instead the
-results of th'e existing engine design and its iinpac t on component life is presented.
3.3.1 Cylinder Heads Cylinder head removal and failure rate are very high. Numerically, l
. 287 heads were removed for corrective maintenance with an average time between removals of 2900 hours0.0336 days <br />0.806 hours <br />0.00479 weeks <br />0.0011 months <br /> as shown in Table 3.lA.
l This equates to every head on both of the engines (32 heads) being changed nine (9) times during their operating life to date. Comparing this to a TDI suggested reconditioning cycle of 8000 hours per head, or approximately four (4) times in 32,000 hours of engine operation, this means that the heads
, have been removed in excess of twice the scheduled maintenance frequency.
i The reason for the head removals were varied, as is shown in Table 3.1B. However, the types of failures could generally be described as being integral to the head and its construction i
and/or reflective of the head materials. The use of cast steel ,
for a head material gives the manufacturer a superior material
! relative to the mechanical and metallurgical properties, particu-larly where the manufacturer uses a welding deposition technique (hard facing) for the valve seats. However, the detrimental
- - 3-7 I
9 fceture of cast steel is the pooror castability .of cast steel (versus cast iron) and the requirement for different and more
. closely controlled foundry casting techniques. The results
)
1
, of some of~ these casting problems have been representative of the types of failures observed in the cylinder heads. Specific failures are head cracking and fire deck warpingfrom high
- stress areas, and porosity from . gas and contaminant inclusions.
l Additional casting technlque problems which have been observed in the heads have been core shifting which has resulted in thin cross-sections and misaligned cooling passages.
l Two additional problems which have plagued the head construction
- and interface areas are the exhaust valve guides and head warping along the 3-9 o' clock axis. In the case of the valve guide
! problems, insufficient documentation was available to reflect the ' number of occurrences chargeable to guide failures or valve guide induced failures' such as carbon build-up on the valve stems which resulted in stuck valves or guide damage. The 1975,
'75 and '76 files contain reports of pieces of valve guides breaking of f and causing foreign object damaged (F0D) to turbo-l chargers but insufficient numerical data has resulted in this l
type of failure being omitted from the historical summary.
i However, the head / valve guide area has been subject to continuous modifications starting with the 1976/77 overhaul when all the j ' guides were machined flush with the exhaust gas passage end i
continuing to the 1982/83 overhaul when additional valve guide length was removed and a valve guide oiling / sealing (Trabon) system was added to control the rocher box _ sooting problem.
i 3-8
,: l _ . - - -- . - . . . _ _ , , .
,9 A_
' [ HEAD ~ BOLTS (8) -
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UNSYMETRICAL gds 13iN'k'
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~'i'.'_'._,__.6._4. g g'., ,'
g -9 p*, SPACING OF WARPING v-
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___ a Figure 3.2 -l DMRV-16-4 Cylinder Head and Valves
t s
Ths- head warping . - problem caused two secondary modes of failure.
One ' mode is the . fire deck warping which has resulted'in internal
- cracking' while. the other is excessive meth1 removal during head reconditioning which has resulted in a shortened head life. _The second mode .o f failure due- to head warping is .the premature unloading of the fire ring gasket which results in
^
the 3-9 o' clock fire ring burn 'out. Reference to Table 3.lB lists 77 head warping / fire ring F-ilures which does not coincide with an observation made by the Chief Engineer (M/V COLUMBIA) where he estimated that M -807. of the heads removed showed 4
- fire ring distress (brown streaking in 3-9 o' clock) or fire ring failure (black streaking in 3-9 o' clock position). TDI has attributed the cause for this type of failure to be the unsymmetrical head bolting pattern around the 3-9 o' clock axis, as illustrated in Figure 3.2. This unsymmetrical pattern results because of the nearness of the ad-joining heads which does not physically allow a head bolt to be placed on a regularly ,
- spaced circumferential bolting pattern. The subsequent bolt
- tightening result's in a bending moment to be formed (or hogging) t i perpendicular to the 3-9 o' clock axis. TDI. has reinforced the interior head area perpendicular to the 3-9 o' clock axis with i
- a " strong-back". Heads with this design modification are presently- ,
in service for a total of approximately 6500 hours0.0752 days <br />1.806 hours <br />0.0107 weeks <br />0.00247 months <br /> with reported failures of three (3) heads in that period. The scrapping of i
twenty (20). heads with an average life of 13,300 hours0.00347 days <br />0.0833 hours <br />4.960317e-4 weeks <br />1.1415e-4 months <br /> represents
! - a- high rate of failure when it is considered that this represents
( 627. of the total heads in service. The reasons for scrapping 4
3-10
..,s_,, a , , , - - .-4wy -,...wy,,m,, ,-,.,g-. .-e py --r, ,-p n--- .--,----r
---.--T--N-t--r-
have been thin fire decks (due to repeated machining cuts during -
reconditioning and/or shifted casting cores) or unrepairable interior cracks or porosities or cracks between valve seats, or non-concentric valve stem to valve seat diameters. Some of the heads have been scrapped during TDI factory reconditioning due to valve bridge cracking during valve seat welding deposition.
The data presented in Table 3.lA indicates a high failure rate for reasons of both design and material selection.
3.3.2 Cylinder Liners The cylinder liner removal and failure rate is very high. The most common reason for liner removal is the necessity for honing to restore liner roundness or surface quality when piston rings were changed. Other reasons fo.r liner removal are attributed to the lower liner to block seal failures which occurred during the first two (2) years of service. Table 3.lA lists 138 liner removals for corrective maintenance with an average time between removals of 5400 hours0.0625 days <br />1.5 hours <br />0.00893 weeks <br />0.00205 months <br />.
The failure mechanism may be attributed to several coincidental factors. The first factor is gauling and scoring of the liners due to embedded materials between the moving surfaces. In some cases this has been from foreign matter, o r. chrome from the ring surfaces that has been spaulled or flaked from the compression ring wearing surface which has become embedded in the piston crown, other rings or the piston skirt with the resultant scoring.
The second factor is the premature wear caused by the unburned 3-11
Figuro 3.3 Pormanont Linor Deformation, Boro Diameter
. no r. __ .. _ . _ . . . . _ _ . __..._I =
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t ri i1 l2 1i tu l '. t6 cy t a nelar siew.ce s , c ioc= w i s,- tr,n ni.nt reont Jacobson Repo ~l r-12 31 March 198
/
- 9- -3
~
A OVALATION
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s b ,a NEW CYLINDER LINER DIMENSIONS I
3-12
Sta rium rel H.ii n Enci .. .
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1 2 3 4 'i 6 7 8 i. 2 3 4 5' 6 7 8 Right Ba nk I.e { t Bank Cy1inder location Legend
-.- .-12 & 6 0 ' C lock 3 & 9 0' Clock
.. --- Ave ra ge Original Diameter 19.000 19.002 I
1 l
Figure 3.4 Bore Diameter, Engine Blocks M/V Columbia j (From Fig. 3 Jon Jacobson Report, March 31, 1981.)
l l
\
l 3-13 I
carbon due to the incomplete combustion in the cylinder. Here, many factors are at work, including raw fuel impingement on the cylinder walls, abrasive carbon wear between the moving surfaces and potential hot spots from the partial combustion process. Improvement of the combustion process should help ,
to extend the cylinder life due to reduced carbon generation and abrasion.
Another persistent type of failure attributing to a premature liner scrapping has occurred as the result of liner ovalation as reported in Reference No. 6 and No. 7 during the 1980/81 overhaul. In this instance liner deformation has been observed
. as a direct result of block def~ormation. That is, the liner ovalates to an increasing dimension in the 6-12 o' clock position (athwart ships) with a decreasing dimension in the 3-9 o' clock position (fore / aft) where the liner is clamped in the counter-bored block lip area. The observed measurements from Reference No.
6 is reproduced in Figure 3.3 to graphically illustrate the observed change in the liner dimensions. The result of this liner deformation is the ensuing ring / piston wear distortion and premature liner wear that occurs as the moving parts try to conform to the dimension char.ge s . The- magnitude of this problem is more graphically presented in Figure 3.4 from Reference No. 6, showing the block deformation which ultimately deforms the liners.
l Of particular importance is the repeatable dimensional change in cylinder number 4 (mid block area) for both engine blocks 3-14
.- - ~ __
- in the- port- and starboard e'ngines. The short
- liner l'ife of 19,600 Lhours, as opposed to..a projected- 50,000-100,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br />,.
l
- could be
- attributed to the high wear rate that has~been accelerated ~
. b y .- incomplete combustion' and by the mechanical forces char
[
'caus'e the cylinder,te ovalate.
- 3.3.3 Pistons The~~' number of pistons removed (149) has been
- influenced by some of Lehe- other- component corrective maintenance actions, such as worn -liners and failed connecting rods. However, the. e-were several. impending failures of bolting mechanisms and ~ crown to skirt oil seals that, upon piston removal, were detected
).
and corrected before a catastrophic casualty occurred. Reference.
to Table 3.1B. lists ninety-two (92) . . piston removals specifically -
for maintenance or modifications to the piston crowns or skirts. -
, The type of modifications made to the piston consisted of decreasing 2
i crown diameter, modifying tube oil passages and seals, and machining modifications to ring grooves, and piston skirts.
l l Modifications of this nature are often considered a product r
j improvement, but in many cases are really . design corrections.
i j Piston crown fastener problems have been observed at various
! intervals. ' Records indicate that' -several crown to skirt bolts i
have broken, or in the case of several overhauls, these same 4
I bolts have been found loosened from the specified torque level.
This problem . continues to manifest. itself by the observation of fretting (metal-to-metal movement an'd wear) under the bolted i . surfaces and bolt washe s. The fact that this occurs indicates. -
that there is surface moven. tnt under high stress conditions.
3-15
Piston rings have required frequent change out and scrapping due to accelerated wear and, in the case of the compression rings, due to chrome overlay chipping or flaking from the surface which has become embedded in the crown, other rings, the tining of the piston skirt or in the liner. Where these flakes have been embedded for some time, deep scoring usually results, as shown in Table 3.lA. One hundred forty-two (142) piston ring sets have been replaced on the thirty-two pistons over the 30,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br /> of running time. This equates to approximately four (4) renewals of rings during a time interval when only one renewal should have been necessary. An improvement in cylinder combustion should decrease this frequency as outlined in Section 3.3.2, " Cylinder Liners".
3.3.4 Master and Link Connecting Rods The articulated connecting rod has been the location of failures which could have been catastrophic if the cracked rods had not been discovered when they were. Prio,r to 1979, there had been one failure by cracking of a connecting rod, coupled with a link rod bearing failure. This same type of failure was later discovered at approximately 21,000 hours of operation in the Fall of 1979. At that time, approximately 257. of the link rod boxes were found fractured in the link pin area between the link pin bushing and the serrated joint. Figure 3.5 shows the orientation of two typical crack propagations (crack A & B) which were found in two (2) separate rod boxes. These fractures occurred in high stress. areas in a high cycle fatigue mode, as illustrated, following.
3-16
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5.'g ,, , ghg
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SECTion A.- A.
Figure 3.5 Master Rod and Connecting Rod Box Assembly
. - ~ -. . - .- . .
Assum2:
Average Engine Speed g = '380 Rev/ Min
.Therefore:
At 21,000 Hours
= 21,000 HRS x 60 Min x 380 Rev/ Min x 1 Cycle M .Rev
= 4.8 x'10 8 Cycles..
Which, by definition, is above low cycle fatigue 10 cycles j ~.and below infinite life where cycles exceed 109 cycles.
- The high. stress ; area was reduced by increasing the cross-sectional-area when TDI decreased the connecting rod bolt size from 1-7/8"-
to 1-1/2" and changed bolt configuration and materials. Additional.
modifications were made to the link rod box external contours by increasing radii to decrease stress concentration areas.
The rod box has exhibited other signs of distress in the ' link pin bushing. This has been addressed .by, a change in bushing j materials. The rod box also houses. part of the connecting- rod g
bearings which have been subject to failure. The 1979/80 engine overhaul disclosed one broken and failed bearing. Upon inspection, j many of the other bearings were observed to be showing signs
[ of distress in the form of fretting and carbonized oil deposits.
Typically, these observations are associated with excessive temperature and/or high loading. Confirmation of this phenomenon was provided by ' Northwest Laboratories II I when a metallurgical examination was made of the No. 6 rod bearing. Table 3.lA lists fifty (50) connecting rod bearing change outs in 30,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br /> 3-18
r of engine operation which represents three (3) complete renewals in a time interval when only one renewal should be expected.
Another area observed to be a point of high loading forces )
is the serrated joint-between master connecting rod and connecting rod box. The serrated surfaces have shown signs of stress in the "V_" of the.serrations in the form of fretting. This phenomenon is illustrative of metal-against-metal movement under high loading conditions. The only corrective action initiated to date to control this problem, is the action taken by the ship's engineers; wherein, upon component tear-down, they will hand
- dress and polish these surfaces to. effect the best bearing surface possible to distribute the loading. It should be noted that the ship's crew typically will improve the surface finish relative to the machined surface "as received" from the factory.
However, even with the care that is exercised by the crew to effect a good load bearing surface, there is still fretting observed upon component disassembly. This would indicate that i relative movement may be induced by either a partial relaxation of the bolting forces due to uneven torquing or vibratory forces i induced from the cylinder firing loads and/or crank.s'aaft. Addi-
] tional supporting evidence whi.ch indicates that a p roblem exists in this area is the fretting and gauling observed between the
] bolt head and washer surfaces , and washer surfaces and connecting rod surfaces. TDI has made washer material changes in an effort to control the fretting; however, subsequent examinations have i
shown that this problem still exists.
3-19
A conclusten which can ba reach:d by the number of failures and the foregoing . discussion is that the articulated rod and its components experience complex and highly loaded surfaces due to the various modes of failure and distress that have been observed in both the structural parts and bearing surfaces.
3.3.5 Camshafts Reference to Table 3.lA lists a total renewal of four camchafts (two (2) per engine) for the engines at 24,000 hours. This numerical figure could be misleading if it is interpreted as a total failure of the camshaft. In this case, a number' of cam lobes were worn beyond acceptable limits and renewal of these lobes was necessary. However, due to the design of the camshaft, the cost of a new shaft was less than the repair cost of the old shaft. Althrough this is a design decision
- made by the manufacturer, it is considered a premature corrective maintenance item relative to the total life expectancy of the component in this application. .
3.3.6 Main Bearings Table 3.lA shows a total bearing replacement at approximately 16,000 hours. The action that was actually taken was to swap the lower main bearing for the top main bearing because the botton main bearing had worn beyond maximum allowable limits.
TDI's preventative maintenance schedule lists 24,000 hours as the first time interval when main bearings should be inspected and replaced, if required. If a comparison is made between other typical medium speed diesels and their anticipated component life, it may be realized that the COLUMBIA's 'ain m bearings experienced premature wear.
3-20
One condition which may have contributed to this is the incomplete combustion experienced with the Elliott turbochargers and- the resultant high carbon loading imposed on the- lube oil. The high carbon loading in the lube oil was further compounded because of the inability of the lube oil system to continuously purify the lube oil and remove the carbon particles. This is a function of the existing tube oil system design where a single purifier is shared between the two engine lube oil sumps on a rotated basis. The addition of another purifier would permit the lube oil systems to have individual dedicated purifiers, filters and hence, continuous contaminant removal for each engine. This would result in better tube oil quality.
3.3.7 Cylinder Block Table 3.lA lists the scrapping of four (4) cylinder blocks.
The reasons for replacing these blocks were based primarily on two basic documented observations. The following narrative is a summary from the " Engine Rebuild Report", by Jon O. Jacobson, March 31, 1981. The first observation was the deformation or lowering of the cylinder liner block counterbore lip, as illus-trated in Figures 3.6 and 3.7. The mechanism by which this was happening is illustrated in Figures 3.S and 3.9, wherein the counterbore lip was cracking under the high stress of the cylinder head hold down force. Non-destructive testing was employed to determine the extent of the c racki .:g in both engines. The results are presented in Figures 3.10 and 3.11. The , magnitude ,
of the cracking, the extent of the cracking, and the potential for the liner " dropping" into the crankcase, with the ensuing catastrophic results, provided a strong case for block renewal.
Figuro 3.6 Upper Cylinder Liner & Block .Section CYLINDER HOLD DOWN FORCE I eg.so, I
' " 9 p COUNTER BORE LIP r '- A r
/8.997 M
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U LINER MOVING INTO CR ANKCASE
- Reprinted from: " Engine Rebuild Report-M.Y. COLUMBIA" Jon O. Jacobson, March 31, 1981 3-22
Figuro 3.7 Cylinder Configuration, Engine Block l;i5
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' Reprinted from: " Engine Rebuild Report-M.V. COLUMBIA" Jon 0. Jacobson, March 31. 1981 3-23
Figuro 3.8 Nondestructive Testing, Cylinder Block, Shear Cracks, Counterbore Lip
'12 NO. Cyl. _2.
Bank: Right Eng: Port i
9 3 4 .
1 6
1 l
' Reprinted fros: " Engine Rebuild Report-M.V. COLUMBIA" Jon 0. Jacobson, March 31, 1981 3-24 9
Figuro 3.9 Nondestructive Testing, ,
Cylinder Block, Delamination Cracks 12 No. CyL : 3 Bank: Left Eng: M l
9 3 p .
1 6
6 r
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' Reprinted from: " Engine Rebuild Report-M.V. COLUMBIA" Jon 0. Jacobson, March 31, 1981 1
3-25 .
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Crack Crack WWWWWWtWQ Crack Crack LEFT BANK Crack ,
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1 RIGHT BANK Figure 3.10 Nondestructive Testing, Port Main Engine
" Reprinted from: " Engine Rebuild Report-M.V. COLUMBIA" Jan 0. Jacobson, March 31, 1981
Dalaminstion Crack Crac i
n n n d '"' n 8 ' '
7 ' l 6 I I 5 ' '
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_LEFT BANK FRONT i
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> 1 6 t 1 5 i I 4 ' i 3 mi 1
2 ' l 1 l Cra Cracks Crack l RIGHT BANK Figure 3.11 Nondestructive Testing, Starboard Main Engine l
- Reprinted from: " Engine Rebuild Report-M.V. COLUMBIA" l Jon 0. Jacobson, March 31, 192; i
i
A second fcctor which contributcd to the ultimmte dacision to replace the cylinder blocks was the continuing cylinder liner counterbore diametral distortion which was maximized at the number four (4) cylinder locations (mid block) on all the blocks. A summary of these measurements was previously presented in Figure 3.4. The significance of this non rsymmetrical dimensional change was the effect it was having on the cylinder bores and the cylinder liners relative to a time base of 24,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br />. If no improvements are made by the manufacturer, it can be predicted that the blocks would have to be repigced at least two (2) more times in the twenty (20) year life of the vessel.
The " Engine Rebuild Report" also investigates the inadequate and irregular block to crankcase bolt torque values documente' the at time of engine overhaul. An additional observation tr made of the fretting or apparent movement which took plac.
between the cylinder block and crankoase base surfaces. Tv.
- problem areas arise here. The first is the implication that correct bolt tensioning was used at the time of manufacture and assembly, and/or that thermal or cyclic loading contributed to the relaxing of the tension which contributed to the relative
- surface movement; or, that improper tensioning occurred at the time of assembly and that the surface fretting was the i result.
i The basic conclusion that can be drawn from the preceding is i
that the block had to be either replaced or repaired due to the dimensional changes and the casting cracking that was l
3-28 I
e
devaloping around tha cylinder lip counterboro. Tha fact that an . acceptable repair was not presented which addressed the cracking problem left AMHS with the only ophion of replacing the blocks. The fact that multiple cracking did occur indicates that the manufacturer has a design problem in this area of the engine.
l 3.3.8 Major Overhauls Reference to Table 3.lA shows four (4) major overhauls at 6,000 hour0 days <br />0 hours <br />0 weeks <br />0 months <br /> intervals during 30,000 hours of engine running time.
A major overhaul was charged against the engines as a Gnit (two engines, four times) any time it was necessary to conduct corrective maintenance, which included liner removal, piston ring replacement and replacement of multiple bearings, either connecting rod, main or articulated link bearings or pins.
Coinciding with this work was the routine turbo cleaning and head reconditioning. Major overhauls occurred during the overhaul periods of 1975/76, 1978/79, 1979/80 and 19,80/81.
An overhaul interval of 6,000 hours represents a rate four (4) times faster than expected between anticipated overhauls for either TDI or other diesel manufacturers.
3.3.9 Turbochargers
- The turbochargers have historically been an item of high main-tenanca with multiple types of failures, including leaking oil / air seals, bearings, nozzles, rotors / cracked casings and fasteners.
Reference to Table 3.lA lists forty (40) removals which coincide with forty (40) overhauts for the four (4) turbochargers on the 3-29 r . . . , . . . - _ _ _ ._
two onginas (two turbos / engine). Tha 2,300 hours0.00347 days <br />0.0833 hours <br />4.960317e-4 weeks <br />1.1415e-4 months <br /> between correctiva maintenance actions means that at least one turbo would require removal sometime before the annual overhaul. Table 3.lA notes a TDI recommended scheduled maintenance cycle of 8,000 hours with other diesel manufacturers listing a TB0 of 12,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br />.
The turbocharger and engine performance with the Elliott Turbo-chargers was covered in Section 2.0. The reason for the recent retrofit of the turbochargers was based on the analysis of the data as previously recorded, including the turbochargers' inability to deliver a high enough quantity of air at an acceptable manifold pressure.
3.4 Summary of Findings It can be seen from the foregoing Section that major moving components of the engine f ail'ed or required an inordinate amount of corrective maintenance at a significantly higher rate as compared to either TDI's recommended scheduled maintenance or other typical diesel manufacturers' TB0s. The types of failures, and number of failures of some of the major components indicates design deficiencies in these components. Two critical components which have been subject to failure, which are not typically expected to routinely fail, were the articulated connecting rods and cylinder blocks.
Another item of significance is the extended overhaul time required due to the greater amount of corrective maintenance which impacted the operational schedule as illustrated in Figure 3.12. The Figure .shows that actual maintenance periods exceeded 3-30
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l pro jeated periods by 307,, resulting in a loss of thirteen (13) mont. a s of potential revenue. Original planning records show generating time that for the five (5) years preceding 1980, maintenance periods of three (3) months were allocated for the entire ship. Post-1980 overhaul periods witnessed a consistent 4
lengthening of the maintenance cycle as dictated by the engines' requirements. Finally, in 1982 and 1983, the maintenance period had increased by a total of one month and the in-service dates were being set to accommodare the engine overhauls.
One item whic.h has not ber addressed previously is the operation of the engines on heavy fuel oil (HFO). Original contract specifi-cations required these engines to opera te on HFO. Demonstration of HFO operation was accomplished upon delivery of the ship and for approximately eight (8) months after that. However, the HFO operation was discontinued after the initial eight (8) months. The records did not disclose why HFO operation was suspended but the following observation can be made relative to what the impact would have been on these engines if HFO had been burned. A survey of other medium speed diesel manufac-
\
l turers' TBO schedules typically reduce length between overhauls by approximately fifty percent ( 507. ) . It can be projected that
- the impact of continuous HFO operation on these engines would have resulted in considerably higher wear and failure rates than those recorded in Table 3.1.
1 l
1 3-32 I
. . \
~4.0 ESTIMATED POTENTIAL REDUCTION OF COMPONENT FAILURES AFTER ENGINE DE-RATING 4.1 Introduction
- M/V '. COLUMBIA had. been operating at reduced power levels, approxi-mately 7,000 HP, for approximately three or four years prior to the 1982/83 overhaul and de-rating. The recent de-rating,-
inclusive of the C-17 turbo modification with a projected engine rating of 6164 HP @ 403 RPM Maximum Continuous Rating (MCR),
represents - an additional 147. reduction of power from the preceding reduced operational power levels. Therefore, the lower pgwer level does not represent a radical change from the previous years' operating scenario. A lower BMEP should _mean lower trans-mitted forces to the various engine components. Likewise, the-replacement turbocharger with greater air. delivery capability should enable the more complete combustion of the fuel thereby reducing the stress and wear rate on combustion related components, such as piston rings and cylinder liners.
4.2 Projected Corrective Maintenance and Expected Component Life ,
The following subsections present a review, and where possible, analyses on the components which would be affected or which were subjects of early failure and replacement, as ' presented
~
in Table 3.1.'
4.2.1 Cylinder Heads The de-rated engine will reduce thermal and mechanical stresses induced by combustion. The improved air flow to the cylinders 4-1
with the resultant towaring of the exhaust gas temperature should .be beneficial to combustion gas path . components such as valves and valve seats. l The installation of the Trabon system wc.s initiated to reduce the carbon / soot loading in the lube oil via the rocker boxes.
The greatest contribution to soot reduction will be the improved combustion offered by the turbocharger modification. However, i based on the premise that the Trabon system (oil injection 4
around the valve stem in the valve guide) will eliminate the
-valve stem / guide blowby, it is in our opinion, of margihal value. This judgment. is based on the continued level of exhaust gas pressure in the . form of dynamic head that results from
- the expansion of gas from the cylinder at exhaust valve opening where typically the gas reaches sonic velocities.
! Head cracking may be lessened due to the reduced thermal stresses
't experienced during operation. However, internal head cracking and porosity leaks and core shifting due to ' manufacturing problems will probably remain at the same level as witnessed by the cooling outlet problems which were experienced with the 16 i
new heads.
Head failures due to warping in the 3-9 o' clock position has l been addressed by the manufacturers by the addition of a " strong back" (reinforcing perpendicular to the 3-9 o' clock axis).
Theoretically, this should contribute to the solution of the 4
problem. A review of the records indicates that in 6,800 hours0.00926 days <br />0.222 hours <br />0.00132 weeks <br />3.044e-4 months <br /> of operation, two (2) heads have. experienced removal because
- 4-2 4
9
of water lecks and one (1) head has baan rcmoved with no reason given. No observations were recorded relative to the fire ring gasket or fire deck ' warping. The average removal ra te is one unit for approximtely every 2,300 hours0.00347 days <br />0.0833 hours <br />4.960317e-4 weeks <br />1.1415e-4 months <br /> of ' opera tion (all heads were removed from the S.M.E. which had the sixteen (16) new heads installed).
4.2.2 Cylinder Liners Cylinder liner removal rate should decrease and life expectancy should increase from improved combustion with the resultant decrease in soot generation and hence, reduction in a bra sive particles. The lacquering problem which has been observed in the liners should also be reduced because of the improved combustion.
Lacquer is actually a combination of resins, soot, oxygenates, oil and water produced by oxidation at combustion temperatures.
The increased presence of soot acts as an increased nucleus site wherein the soot precipitates on the cooler liner walls and the resin-like substance concentrates around and between the soot particles. Reduction of the soot particles should therefore reduce the lacquer accumulation.
Liner removals influenced by the dimensional change of the cylin'der block are not expected to change due to the mechanical deformations imposed on the liners by the block. Therefore, it can be projected that some liners will reach the end of their useful life at approximately 20,000 to 30,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br />.
l 4-3'
4.2.3 Pistons Piston removals for ring replacement should decrease for the same related combustion improvement reasons discussed in the cylinder liner Section 4.2.2. However, the same dimensional change problems which affected the liner roundness will also cdversely affect the ring wear rate and life expectancy.
Piston removals for modifications to crowns, skirts and fastening mechanisms should be finalized as this model of engine has been in existence for approximately ten years and therefore the manufacturer should have incorporated, by now, all the related design corrections gained through experience on existing engines. '
4.2.4 Master Link and Connecting Rods The articulated rod removal rate frequency can be expected to decrease due to the decreased loading on that assembly with a subsequent anticipated increase in component life. However, the approximate 14% reduction in engine de-rating is not a significant reduction when the previous operational power levels of 7,000 HP are considered and therefore improvement may be marginal.
The link rod bearing and pin useful life expectancy should increase due to the overall decreased loadings. However, this whole assembly is considered a highly loaded part relative to the dynamic forces induced on ther" components. Past corrective mainten-ance procedures have been concerned with failures induced by poor quality control of the pin and bushing. If these manufacturing problems are resolved, then the remaining problem of bearing wear should be minimized.
4-4 L_ _
1
. i ,1 Tha- fosteners and link rod box cracking phenomenon was addressed by) TDI as,lpreviously discussed in, Section 3.3.4. The decrease in bolt diamete'r from 1 - 7 / 8 ',' to 1-1/2" resulted in-a net width gain of .75 inches on the total width of the connecting rod essembly. This modificaticn plus the radius changes should increase the expected life of th6 component. It should be realized, however, that the shorter connecting rod for an equal stroke end crank throw radius has the greater angular swing and greater side thrust, hence greater loading on the piston and link pin and bushing. Intergral with the fastening mechanism is the serrated joint which has been the ' site of repeated observations of fretting and therefore relative movement. The employment of the serrated joint is recognized for its value in transferring stress loads in the plane perpendicular to the serrations, however, it would appear that the resultant forces acting normal to the serration plane are contributing to the induced movement.
Stress Transfer Plane 1
(*
NORMAL < FORCES
= >
< =
In a four-cycle engine this is especially pronounced when it is_ considered that there is a rede'rsaU'of loading on each connect-ing ' link rod from compression to tension on each cycle along the legs of the "V" which is formed by the centerline of the connecting rod axis. The serrated joint surface then is constantly 4-5 '
s
b:ing subjected to the normal and parallel forces induced by the two pistons. This type of a joint is also subject to con-troversy among manuf acturers because of the argument that bearing distortion is more easily induced because of variation in serrated surface . irregularities and bolt tightening. Two of' these factors have been observed to date. This is further reinforced by the failure mechanisms noted in the Northwest Laboratories Report (Reference No. 17) which concluded that the connecting rod bearing was damaged by localized over heating.
This finding reinforces the observations made during the 1979/80 overhaul of distressed connecting rod bearings with localized loading spots, fretting damage and carbonized oil deposits.
l 4.2.5 Camshafts The camshaft wear rate should be decreased slightly due to the reduction of carbon in the lube oil. However, because the cam shaft is unaffected by engine de-rating except for the lower engine speeds, it is expected that the ' useful component life will remain approximately the same.
4.2.6 Main Bearings The main bearing wear rate should be decreased slightly due to the reduction of carbon in the lube oil and the slightly lower crankshaft loads due to lower BMEP. It is our opinion that the bearing will experience approximately the same 16,000-20,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br /> before requiring renewal.
4.2.7 Cylinder Block l
l 'The engine de-rating will lessen the thermal stresses due to 4-6
the reduced heat rejection to the jacket cooling water and will lessen the mechanical stresses due to reduced BMEP.
The two major items of concern are the cylinder liner counterbore dimensis al symmetry and reoccurrence of counterbore lip cracking.
In the instance of counterbore symmetry, we would expect a reoccurrence of this phenomenon unless structural and design changes have been made to the cylinder blocks. In the case of the cracked cylinder liner counterbore lips, we would expect this to reoccur unless design changes are made to the block in that area. In the absence of formal notification by ,TDI that a change has been made and is the subject of a retrofit, we would anticipate block lip cracking in the future. It should be noted that unsymmetrical oval counterbores can be repaired if that is the only problem with the block at that time. However, a structural weakness with the potential for catostrophic failure should still be considered as a replacement item.
The cylinder block to engine base fretting that was observed during the 1980/81 overhaul has also been observed in other manufacturers engines. The relaxed cylinder block to engine base tie rods contributed to the severity of the fretting and as Mr. Jacobson recommended, this can .be minimized by periodic retorquing of the tie rods as availability dictates. If cylinder block to engine base fretting does occur, this may be repaired by various resurfacing techniques. It should be realized that once fretting occurs to the extent that a minute relaxation of ten-sioning occurs, the fretting effect accelerates at an increasing 4-7
rate ' with - the subsequent increasing relaxation of the tie rod tensioning.
4.2.8 Major Overhauls The total effect of engine de-rating will result in an improved level of component quality where those components are exposed to the combustion gas path. However, we do not feel- that a quantum improvement will result which will approach the TDI suggested maintenance time intervals.
4 4.2.9 Turbochargers There is no turbocharger history to draw upon to s pecul'a te i
on a corrective maintenance interval; however, the TDI suggested maintenance schedule lists 8,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br />.as the TB0 cycle.
There are several comments that are germane to the turbocharger and exhaust system. The C-17 turbos do not have' the capability l for localized blocking or otor restraint devices in the event of failure. This is considered a drawback because the entire engine must be shutdown in the event of a failure. Most turbo manufacturers have a method for blocking the rotor so that the engine may be run naturally aspirated at reduced power.
The second item of concern is the cantilevered mounting method which was used to intergrate the turbos into the engine exhaust i
and manifold air systems. A cursory examination of the foundation
~
and bracing assembly would raise doubts as to the long term integrity of the structure due to the amount of weight which is cantilevered over the front of the engines. The weakest 4-8 I w - - e- _v,--w
point of the assembly appears to be the point of attachment at the engine block.
The final comment concerns the exhaust system which has been the object of repeated repairs, replacements, and design configura-tions. Historically, this has been an area of high maintenance.
It was not reviewed in Section 3.0 because it is n o t- one of the moving parts within the engine. Therefore, this system should be subject to less stress due to the lowered exhaust temperatures and hence, require less corrective maintenance.
However, based on historical data, we would still expect to see this system receive a higher than usual amount of service.
4.3 Additional Modifications and Corrections of Problems Created By Engine De-Rating 4.3.1 Lube Oil Systems The lube oil systems current configuration utilizes a common purifier which is switched between engines; however, each engine.
is equipped with a full flow filter fqr constant filtering capability. A change should be made in the lube oil system so that another purifier is procured to effect a dedicated lube oil system. The inclusion of a dedicated system would significantly improve the lube oil quality and reduce the carbon
. loading on the system. The piping modifications to the lube oil system should definitely include the repair / replacement of the duplex full flow lube filter diverter vain to permit change '
over during operation instead of the present requirement to shut the engine down.
4-9
An optional addition to the lube oil system would be the installa-tion of a polishing system which would take suction from the i 1
lube oil sump and return to the sump.
4.3.2. Cooling Water System Recent modifications to the cooling water system have ,resulted in .the inability to utilize the stand-by jacket water cooling pump. Failure of this pump in the present configuration would require engine shutdown. The stand-by jacket water pump is required to ensure continuous engine operation.
The temperature control system may also require modification to ensure sufficient temperature in the jacket water system and turbocharger after cooler. Initial voyage results indicate that a deficiency exists in this system.
4.3.3 Turbochargers In Section 2.0 it was shown that the turbochargers were not matched to the engines. It is impe ra tive , therefore, that a proper match be made of turbocharger to engine to ensure future engine reliability and efficient operation.
4.3.4 Waste Heat Boiler Subsequent sea trial feedback has indicated that insufficient i
exhaust gas is being routed through the waste heat boiler with the consequential requirement that the auxiliary boiler make up the difference in steam shortfall at increased cost due to the ad--
ditional fuel consumption of the boiler. This will require 4-10
the installation of an equalizer baffle or orifice to ensure su'fficient heat to the boiler.
4.3.5 Engine Performance Optimization The sea trial data indicates that engine power output is lower than what is required by contract and that the engine is not at. its most efficient de-rated operating level. Additional problems are seen with the lower than expected firing pressures.
Problems. of this nature are usually associated with fuel timing and metering systems.
(:
l j
4 l
5 4-11 4
5.0 RE-ENGINING ECONOMIC ANALYSIS AND COMPARISON OF HISTORICAL MAIN ENGINE OPERATING COSTS AND EXPECTED DE-RATED ENGINE OPERATING COSTS The cost analyses presented in this section deal with three (3) major subject areas. The first is the quantification of an average annual main engine maincenance and repair cost and the estimated reduction in this expenditure which can reasonably be expected to result from derating. Also addressed are other capital expenditures which, in Seaworthy's opinion, must be made to ensure the reliability and efficient performance of the de-rated engines or which must be made as a result of additional problems resulting from the de-rating project based on the status and results of this effort to date. The third and final area of discussion presented in this section is an economic trade-off analysis which compares continued operation of the de-rated engines , including the additional capital expen-ditures required for reliable and efficient operation, against the estimated cost associated with the re,-engining of the M.V.
COLUMBIA.
5.1 Historical Main Engine Related Cost Review and Development As a basis for establishing an estimated main engine maintenance
-and repair average annual cost, various operating cost records were reviewed, dating from 1976 through 1982. These . included ,
for the most part, purchase order type documents, major main engine overhaul cost breakdown reports and AMHS Fiscal Year 1 1
Expense and Revenue Statements for the COLdMBIA. Costs associated with the current de-rating project were not included. The bulk 5-1
of thase records provided a gross or macro view of engine / engine
. department costs for each year. Because the majority of these records lacked a detailed itemized breakdown .into such areas as individual labor category, spare parts by component, cosumables, contractor or repair facility which clearly identified the associated expenditure as being main engine related, the following approach was taken. Fifty (50) percent of all cost obtained and identified . as accruing during each annual propulsion plant /
engine room overhaul period when the vessel was out of service, were assumed to represent that portion of the total annual j power plant overhaul period costs directly related to Inain engine maintenance and repair. Taking a similar approach for the operating portion (and associated costs) of each year,
, twenty (20) percent of the identical cost categories were taken as being representative of main engine related maintenance and repair costs wh'le the vessel was in service. The cost categories and typical associated elements are listed below.
OVERHAUL PERIOD OPERATING SEASON
- 1. Labor: 1. Labor:
- Base Wages - Base Wages
- Overtime - Overtime
- 2. Commodities: 2. Commodities:
- Spare Parts - Consumables less fuel and lube oil
- Consumables - Spare Parts
- 3. Contractual: 3. Contractual:
- Shipyard - Riding crews
- Service Reps. - Service Reps.
- Other contractors - Others
, 5-2
Equipment:
- 4. Equipment: 4.
- Tools - Tools Table 5.1 presents the estimated annual operating season and overhaul period costs derived for main engine maintenance and repair. .
TABLE 5-1
SUMMARY
OF ESTIMATED ANNUAL MAIN ENGINE RELATED M&R COST, 1976 to 1982 l
Operating Overhaul Yearly Year Season,$ Period, $ Total, $
1976 $ 242,041 $ 87,710 $ 329,751 1977 34,433 64,540 98,973 1978 218,935 242,221 461,156 1979 155,674 449,048 604,722
, 1980 179,994 344,055 524,049 1981 169,184 433,546 602,730 1982 201,598 269,228 470,826 Estimated Average Annurl Main Engine M&R Cost:
$441,740/ Year, Historical Discounting present problems associated with the engine de-rating project, it is reasonable to anticipate that the average annual main engine maintenance and repair costs shown above, t
after resolution of the current difficulties, would be reduced.
While the exact value of this expenditure reduction requires considerable speculation, it is felt tha: an improvement- of 25% is a reasonable approximation. This is based primarily on Seaworthy's past experience in performing similar analyses and a correlation of the historical main engine maintenance V
5-3 l
.and repair _ data including component failure analysis, overhaul reports, _ maintenance and repair related cc:ts and ABS surveyor reports as summarized in Sections 3.0 and 4.0 of this report.
Also taken into consideration was TDI's performance record relative to providing cost effective, permanent and sound engineer-4 ing_ solutions ~ to numerous design, production and (to a lesser extent) operating based engine component failures which have significantly increased this annual expenditure. Specifically, it is- believed that this reduction in M&R costs will accrue from engine de-rating and_ new turbos as a result of minimal improvements in component life and . time between repair and/or overhaul for the following components based on the discussions r
presented in Section 4.0.
, 1. Cylinder Heads
- 2. Cylinder Liners
- 3. Piston Rings
- 4. Articulated Connecting Rod Assembly
- 5. Main Bearings
- 6. Exhaust Manifold / Cylinder Head Jumpers
,. 7. Lube Oil Life (Carbon loading reduction) t 5.2 Propulsion System Modifications Required in Addition To or As A Result of Main Engine De-Rating l As part of the workscope which is addressed by this report an evaluation as to the adequacy of component and systems modifications made as a result of the engine de-rating was conducted. The intent of this investigation was to identify 5-4
and- quantify, in- terms of time and cost, additional work felt necessary to ensure the future operating reliability and efficiency ,
i 1
of the COLUMBIA's de-rated propulsion plant. Additional work, i i
some of which is major, has also been identified and quantified as a result of the performance of the main engines during the ,
March 24-25, 1983 sea trials and subsequent voyages. These modifications, along with supporting rationale, estimates of the time required to accomplish them relative to the scope of the work and Rough Order of Magnitude, ROM, cost estimates for each are provided in Table 5.2. The ROM cost estimates include components / hardware and necessary installation materi'als and labor.
5.3 Re-Engining Economic Trade-Off Analysis An economic comparison has been made which evaluates the continued operation of the existing main engines after de-rating versus the installation and operation of new Heavy Fuel Oil (HFO) capable engines identical to the types specified in Reference No. 2, based on varying values of the assumed remaining useful life of the vessel. The various cost elements, methodologies applied and results are presented in the following para-graphs.
5.3.1 Cost Elements The cost elements established for this analysis have been categorized in two (2) main areas, that of acquisition costs and annual operating costs.
Acquisition Costs: An associated capital expenditure for each 5-5
TABLE 5.2 ADDITIONAL PROPULSION SYSTEM MODIFICATIONS REQUIRED FOR M.V. COLUMBIA
, AFTER DE-RATING AS OF APRIL 1, 1983 Documentation / Supporting Time To Rough Order of Modification / Alteration Rational Complete Magnitude (ROM) Cost
- 1. Lubricating Oil System Installation of 2nd L.O. purifier Overhaul' $150,000 to provide simultaneous L.O. puri- Period fication for both main. engines, polishing filter for.each engine and modification of existing filter valving for improved operation.
- 2. Combustion Improvements It is anticipated that actions in- Operating ~$60,000 cluding F.O. injection timing ad- Season (No vancement, cylinder firing pressure loss of balancing and fuel metering compon- service) ents (pumps, injector, nozzles) may have to be modified / replaced to re-store original design fuel consump-
, tion at the de-rated output.
- 3. Cooling System Modification of engine cooling loops Operating $30,000 to increase jacket water temperature Season (No as part ot' de-rating ptocess and to loss of restore jacket water / fresh water service) pump redundant service capability.
- 4. Exhaust Gas Pyrometer Replacement of existing exhaust gas Overhaul $75,000 System pyrometer system with a more accur- Period ate, reliable and useful system.
e
TABLE 5.2 CONTINUED Documentation / Supporting Time To Rough Order of Modification / Alteration Rational Complete Magnitude (ROM) Cost
- 5. Turbochargers Based on data and discussions pre- Overhaul $470,000 sented in Section 2.0, the turbos Period (Price includes: 4 are not matched to de-rated engine turbos, spares, load profile. Replacement with prop- tools, spare rotating erly matched units is felt to be the element, new tacho-most prudent and reliable fix for meters, transisition this program. ducting, foundation
' engineering & labor)
- 6. Waste Heat Decovered As a result of de-rating there is Operating $20,000 Steam Generoting a short fall in steam available Season (No System for auxiliary and hotel loads. loss of
" service)
This has cuased a noticable in-d crease in fuel consumption as a .
result of continuous operation of the auxiliary oil-fired boiler.
This situation may be rectified by diverting more exhaust gas away from the silencer and into the .
waste heat boiler on each engine.
- 7. Control System Pitch sche'dule and load control poc- Operating $30,000 tion of the main engine control sys- Season (No tem has not been properly set up for loss of the new de-rated engine operating service) profile.
- 8. Structural Leaks in the compressor discharge / Ope ra t ing- $20,000 manifold inlet transition pieces Season (No have been noted and can be expected loss of to increase. Installation of flexi- service) able transition pieces would relieve this situation.
TOTAL $855,000.00
alternative, continued operation of the de-rated main engines and re-engining of the COLUMBIA with HFO capable diesels, was established. These values were assumed to include costs for purchase, installation labor, installation materials, rip-out and other typical activities associated with this type of work.
For the continued operation of the main engines, an acquisition cost of $855,000.00, established in paragraph 5.2 was utilized.
Values of $6, 7, 8 and 9 million dollars have been assumed as a range of acquisition costs, representative of a potential re-engining cost spread for the COLUMBIA, in order to test the sensitivity of the analysis to this potential variable as described in paragraph 5.3.3.
Annual Operating Costs: Because this evaluation is limited in its consideration of only operating side economics, the annual operating expenditures considered were those felt to be directly attributable to main engine operation, fuel oil consumption, lubricating oil consumption 'and maintenance and repair. While new engines might arguably increase the operating seasons for the vessel over continued operation of the existing units as a result of improved reliability and reduced maintenance, the impact of this possibility was not factored as it implies revenue-side analysis which was beyond the scope of this evalua-tion. In deriving the opera' ting cost elements for each option, an optimum operating year of 5400 hours0.0625 days <br />1.5 hours <br />0.00893 weeks <br />0.00205 months <br /> at an average output i of 10,500 BHP was assumed for both alternatives so that the analysis could be conducted on an equivalent basis.
I i
5-8
Fuel Cost: Annual fuel costs were computed for each alternative based on at-sea operation for 5400 hours per- year at 10,500 BHP. For each alternative quoted, manufacturer's fuel rates were utilized and adjusted in the following manner. The existing DMRV-16-4 engines' design quoted fuel rate was increased by 10% to account for a 3% guarantee margin in addition to a 7% increase which is felt to be representative of the deterioration in performance that has occurred based on the historical failure analyses conducted in Section 3.0 and 4.0. A final upward adjust-ment was made to account for the difference in typical Marine Diesel Oil (MDO) heating value versus the heating value ' of Marine Gas Oil (MGO) cn which design quoted fuel rate is based.
Future operation on Heavy Fuel Oil was not considered as a viable alternative for the existing engines based on the documented poor past performance and reliability experienced with the engines while operating on MDO. The new engines' design quoted fuel rate was adjusted in an identical fashion to that described previously including a quoted 5% guarantee margin and an adjustment
, for heating value differences for MD0 and operation on HFO.
The HF0 to be utilized was assumed as 180 CST (1500 second Redwood No. 1, SR1). No increase in fuel rate for the new engines was assumed. Fuel pricing utilized per metric ton was that posted during March 1983 at the Port of Seattle for MD0 and 180 CST fuels. Sample consumption fuel calculations are contained in Appendix 1.
Lube 011: Lubricating oil cost were derived utilizing manufacturers 5-9
A quoted lube oil rates of one (1) gallon /5000 BHP hours for the Enterprise engine and 1.5 grams / BHP-HR for the new engine.
For operation on MD0 a lube oil with a total base number of TBN-10 was utilized due to the low sulfur content in this fuel.
For new engine operation on HFO a lube oil with a TBN of 30 was utilized as a result- of the increased sulfur content of HFO. The TBN designation basically is an indicator of a higher.
content of various chemical'addatives put in the oil to neutralize the potential of increased acid corrosion attack of engine internals when operating on HFO. Lube oil prices utilized were those posted in Seattle as of March 1983. Sample lube oil c6sts calculations are, contained in Appendix ! .
Maintenance and Repair Costs: The annual maintenance and repair costs utilized for the existing engines was the historic rate derived in paragraph 5.1 and adjusted downward, based on future improvements expected from the engine de-rating. Costs for the .new engine were derived utilizing Figure 5.1, a curve of maintenance costs for typical medium speed diesel engines in S/HP-Yr versus fuel oil viscosity. This curve was initially generated by Seaworthy as a result of two research projects performed for the U.S. Maritime Administration dealing with the influence of fuel quality on the maintenance and repair of marine diesel engines and has been ' updated on a frequent basis in published papers and presentations given by Seaworthy personnel. It is felt to reasonably account for the well-documented and ' universally accepted fact that engine M&R costs increase as the' quality of fuel supplied (using increasing viscosity l
l 5 , - . - - e
Figure 5.1 Impact of Varying Fuel Quality on Engine Maintenance, Total Spares, Consurnables, and Labor 25 -
/
20 i : '
/
l 7 0-
- /
f '
n -
" % /
Q. _- /
I -
~
/
i
$ 15 -
_/ --.-
. g -
.J '
s O - '
? o : -
1; 10 -
5 _
0- - l 10 20 50 100 200 500 1000 2000 5000 10,000 REDWOOD # 1 at 100 F
as an indicator) decreases. Appendix ! contains calculations of estimated annual maintenance and repair costs for the new engines for MD0 and HFO operation. The components of the cost derived utilizing this curve are essentially identical to those utilized in developing the historical cost for the existing engines; labor, spare pa,rts, consumables, tools, etc.
Table 5.3 and 5.4 present unit costs utilized for fuel and lube oil in the analysis and a summary of acquisition and first year annual operating costs, respectively.
5.3.2 Economic Analysis Methodology The approach taken in establishing the potential economic benefit associated with re-engining of the COLUMBIA versus continued operation of the existing engines is best summarized in the following manner. It is the determination of whether the annual operating cost differentials (existing engine less new engine annual costs) justify the initial non-recurring acquisition cost differential (re-engining less de-rating modification costa) over the anticipated remaining useful vessel operating life as determined by utilizing the annual cash flow dif ferentials to calculate the following financial indicators: Net Present Value, NPV, Internal Rate of Return, IRR, Simple Payback, SPB and total Life Cycle Costs, LCC. These computations have been performed by micro-computer. The actual computer output data is contained in Appendix H . These and other terms, as applied in this analysis, are defined in Table 5.5.
To take into account the influence of inflation over the investment l
l l
5-12
l TABLE 5.3 .
FUEL AND LUBE OIL UNIT COSTS l 1
Fuel & Lube Oil Types Unit Price
- 1. MD0 $276/ Metric Ton
- 2. HF0 (180 CST) $181/ Metric Ton
- 3. TBN-10 Lube Oil $3.86/ Gallon
- 4. TBN-30 Lube Oil $4.26/ Gallon TABLE 5.4
SUMMARY
OF ACQUISITION AND FIRST YEAR ANNUAL OPERATING COST ESTIMATES Cost Category Existing Engines New Engines
- 1. Acquisition Costs: $ 855,000.00 $6,7,8 & 9 Million (See Paragraph 5.3.;
- 2. Annual Operating Cost:
Fuel Oil:
't MD0 $2,666,440.00 $2,416,380.00 HF0 ----
$1,655,970.00 Lube Oil:
MD0 $60,630.00 $100,090.00 HF0 ----
$122,600.00 Maintenance & Repair:
MD0 $331,310.00 $120,750.00
- HF0 ----
$218,400.00 L 5-13 l
1 TABLE 5.5 I DEFINITION OF ECONOMIC ANALYSIS TERMINOLOGY i
TERM DEFINITION
- l. Acquisition Cost Total value in dollars of all cost associated with the acquisition and installation /modifica-tion of each alternative.
- 2. Investment Period Period of time in years over which the vessel is (Remaining Vessel expected to operate and produce the anticipated j Life) savings, normally, the remaining useful vessel life after conversion or upgrading.
- 3. Method of Financing Source of capital to cover the associated acqui-sition cost, assumed here to be 100% equity by the State of Alaska. (Other sources may Lnclude external financing or combinations of part equity and part external financing).
- 4. Discount Rate The minimum rate selected by an organization
{
which a prospective investment must return, j assumed here as 10%.
- 5. Fuel Price Escalation Rate The annual rate in percent at which fuel price l is estimated to increase throughout the remain-ing vessel life.
- 6. General Economic The annual rate in percent at which the cost of Inflation Rate non-fuel related goods and services is antici- !
paced to increase throughout the remaining vessel >
life. '
- 7. Salvage Value An estimate of the maghet value of machirery com-ponents associated with the conversion or up-grading at the end of the remaining vessel life.
In most instances the only real future value of this equipment is that of scrap which is usually quite small in comparison to the original acqui-sition cost (assumed here as 0.)
- 8. Annual Operating Cost Elements For each propulsion plant alternative considered, a first year's operating cost must be quantified.
Three elements have been assumed to make up the total annual operating costs associated with each alternative:
- 1. Fuel Costs
- 2. Lubricating Oil Costs
- 3. Maintenance and Repair Costs 5-14 i-
TABLE 5.5 CONTINUED TERMS DEFINITION
- 9. Net Present Values, The total value in today's dollars of all future NPV
+
annual cash flow differentials discounted back at the discount rate selected.
- 10. Internal Rate of The rate of interest yielded when the future Return, IRR values of all annual cash flow differentials are assumed to be invested so as to equal the acqui-sition cost differential for the alternative con-sidered.
- 11. Simple Payback,_SPB The break-even point of the investment in years when the future values of accrued annual cash flow differentials equals the acquisition cost differential for the alternative considered.
- 12. Life Cycle Cost, The total projected cost of an alternative over LCC its expected investment life, including acqui-sition and annual operating costs.
4 5-15
' life (rcmaining vessel life), the computational procedure escalated projected annual operating expenditures at a rate of 6% per year *f or maintenance and repair costs and 8% per year for fuel and tube oil costs. This is based on recent historical trends which indicate that annual escalation rates for petroleum related products and by-products have grown at a more rapid pace than non-petroleum based goods and services which would emcompass maintenance and repair cost components. The investment periods utilized here were assumed to be a range of values, from ten to twenty years , equating to the life remaining for the COLUMBIA.
A starting date of June 1, 1984 was assumed, which would allow for time required to re-engine if such a decision were made at this time. Because pricing and cost data was estimated in 1983 dollars, these elements were escalated at previously mentioned rates of inflation to reflect costs as of the theoretical June 1, 1984 start date. These and other pertinent economic analysis input data are summarized in Table 5.6.
TABLE 5.6 ECONOMIC ANALYSIS IN-PUT DATA AND ASSUMPTIONS
- 1. Investment Period 10,15 and 20 years (Remaining vessel life) (See Paragraph 5.3.3)
- 2. Method of-Financing 100% Equity
- 3. Discount Rate 10%
- 4. Escalation Rates-Maintenance & Repair 6%/ Year Fuel & Lube Oil 8%/ Year
- 5. Salvage Value 5-16
5.3.3 Sensitivity Analysis
.To provide. a .
broader scope for the previous described economic analysis, the sensitivity of decision to re-engine versus continued.
4 operation of the existing engines was tested relative to the impact of- varying investment period / remaining . vessel life and acquisition cost estimates for the re-engining -alternative.
, These sensitivity analyses were performed as an integral part 1
of the computer calculations referenced- earlier, the results of which are contained in Appendix 1.
5.4 Discussion of Results i The results of the computer-based economic analysis calculations are presented in Tables 5.7 and 5.8, for both alternatives operating on MD0 and for the existing engines running on MD0 l and the new engines on HF0, respectively. Addressing Table 5.7, first, it can be seen that, for the re-engining alternative i
operating on MDO, i .t is only when the acquisition cost is assumed to be $6,000,000 and the investment period ten (10) years that i .
all financial indicators support the decision to re-engine.
Any increase beyond $6,000,000 as an assumed acquisition cost
! while maintaining the investment period of time available to recover the capital expenditure required for engine change out at ten (10) years, results in various indicators failing
- to support re-engining. These are manifested as a decrease in IRR to zero, or as shown, "No Return" which indicates that the investment would not. be recouped in the basis of annual savings over continued operation of the existing engine within e-9 0
2 5-17
TABLE 5.7
SUMMARY
OF RE-ENGIllltiG ECONOMIC ANALYSIS RESULTS FOR NEW ENGINE OPERATION ON MD0 ASSUMED RE-ENGIllIllG ASSUMED EMAINING WSSEL WE ACQUISITI0t1 COST 10 YEARS 15 YEARS 20 YEARS NPV: - $ 1,2 5 5,124 NPV: $331,499 NPV: $1,722,831 IRR: 4.5% IRR: 10.97. IRR: 13.5%
$6,000,000 LCC o- $1,531,916 LCC 6-$7,031,500 LCC 6 -$14,799,070 SPB: 8.25 Years SPB: 8.25 Years SPB: 8.25 Years NPV: - $2,2 55,124 NPV: - $ 668,500 NPV: $722,831 IRR: 1.4% IRR: 8.3% IRR: 11.3%
$7,000,000 LCC 6-$531,916 LCC 6-$6,031,500 LCC 6-$13,799,070
, SPB: 9.4 Years SPB: 9.4 Years SPB: 9.4 Years NPV: - $3,2 55,124 NPV: - $1,668,501 N PV.: - $277,169 IRR: No Return IRR: 6.3% IRR: 9.55%
$8,000,000 LCC /; $478,054 LCC o-$5,031,500 LCC o-$12,799,070 SPB: 10.5 Years SPB: 10.5 Years SPB: 10.5 Years NPV: - $4,255,124 11PV : - $2,668,501 NPV: - $1,277,169 IRR: No Return IRR: 4.7% IRR: 8.1%
$9,000,000 LCC o $1,468,084 LCC o-$4,031,500 LCC 6 -$ 11,799,070 SPB: 11.5 Years SPB: 11. 5- Yea rs SPB: 11.5 Years
L BLE 5.8
SUMMARY
OF RE-ENGINING ECONOMIC ANALYSIS RESULTS FOR NEW ENGINE OPERATION ON HFO ASSUMED RE-ENGINING ASSUMED REMAINING ESSEl. l.IFE ACQUISITION COST 10 YEAR , 15 YEAR 20 YEAR NPV: $5,110,098 NPV: $9,516,611 NPV: $13,506,697 IRR: 27.27. IRR: 30.47. IRR: 31.37.
$6,000,000 LCCa -$12,554,115 LCCa -$27,847,131 LCCA -550,149,680 SPB: 3.75 Years SPB: 3.75 Years SPB: 3.75 Years NPV: $4,110,098 NPV: $5,516,611 NPV: $12,506,697 IRR: 22.17. IRR: 25.97. IRR: 27.17.
$7,000,000 LCCa -$11,544,115 LCCo -$26,847,131 LCC A -$49,149,680 y SPB: 4.4 Years SPB: 4.4 Years SPB: 4.4 Years 5 '
NPV: $3,110,098 NPV: $7,516,611 NPV: $18,173,831 IRR: 18.27. IRR: 22.57. IRR: 24.0T
$8,000,000 LCCL. -$10,554,115 LCCA -$25,847,131 LCCA -$48,149,680 SPB: 5.0 Years SPB: 5.0 Years SPB: 5.0 Years NPV: $2,110,098 NPV: $6,516,611 NPV: $10,506,697 IRR: 15.0 7. IRR: 19.97. IRR: 21.67.
$9,000,000 LCC 6 - $9,554,115 LCC /_ - $ 24,84 7 ,131 LCCA -$47,149,680 SPB: 5.6 Years SPB: 5.6 Years SPB: 5.6 Years 8
2
the given investmnnt period. Also, a change in LCCO from negntive to a positive dollar value for a given combination of new engine acquisition cost and remaining vessel life does not support re-engining. This is because LCC o has been set up mathematically to- equal the remaining dollar value when total life cycle costs associated with continued operation of the existing engines .
is subtracted from the equivalent costs associated with re-engining.
Thus, as long as life cycle costs for the re-engining alternative are less than those for continued Enterprise engine operation, the LCC o will be negative. A change to positive indicates that this life cycle cost relationship has been reversed.
Also, between the three (3) financial indicators, NPV, IRR and LCC o , for certain acquisition cost / time scenarios in Table 5.7, an apparent conflict seems to occur, that being the fact that the IRR and LCC o values tend to suport re-engining while the NPV value shown is negative. This is due to the fact that IRR and LCC a are computed on the basis of future inflated cash flow differentials while NPV represents the sum of all cash flow differentials during the investment period in todays dollars discounted back at 10%. Thus, a negative NPV may not in itself mean that the re-engining should not be undertaken, but that i
it does not begin to payback until late in the investment period in te r'ms of accuring positive annual cash flow differentials.
This relationship can be easily determined by inspection of predicted annual cash flow data for che alternative in question contained in Appendix 1. Simple payback, SPB, speaks for itself 5-20
l
-in that the shorter this period is in years, generally, the more attractive the investment. The results shown in both Tables 5.7 and 5.8 are' absolute but do not show a clear optimum scenario in that a minimum time period required to recoup the investment or that a certain rate of return has been obtained. These criteria, it is assumed, will be established and factored by AMHS.
Referring to Table 5.8, .it becomes immediately obvious that re-engining for HFO operation is an economically superior alterna-tive to re-engining for continued MDO operation. All that remains to be identified is a satisfactory rate of return on investment and how long in years this return should take, and a ran'ge of satisfactory capital cost / investment period scenarios for re-engining for HFO operation can be selected.
5-21
h .
6.0 CONCLUSION
S AND RECOMMENDATIONS 3
' Based .on the results of the detailed freview and analysis of ;
/
g current performance data and main engine maintenance and repair history and related cost information as presented in previous report sections , . numerous conclusions $nd, resultant recommendations -
have been made which are presented in this section of the report.
For the-- sake of organizational clarity and brevity, the attendant conclusions and recommendations have been divided into the following relevan,t categories:
, s Sea Trial Performance Adequacy of the Engine De-Rating Additional Modifications Economic Evaluation of Re-engining 5
- Sea Trial Performance:
9 !
- 1. The engines as de-rated / by TDI failed to develop the required power outputs as' specif,ted in the work scope of the contract authorizing this work.
a a, 2.
The turbochargers, as it. dica ted by surge problems observed during the trials and on subsequent voyages are not . properly matched to the new de-rated engine operating profile. Emperical data presented in Section 2.0 further supports'this conclusion.
- 3. Numerous other problems of a smaller magnitude also identified in Section 2.0, have developed as a . result
)
+
6-1 P
l of the de-rating work and for the_ most part are un-resolved.
4 ~. Adequate air flow appears to have been provided to by.
the engines the new turbochargers. Brake Mean Effective Pressures- at the new operating -outputs are equal to, or less than, those specified- in the de-rating contract.
- 5. It is possible that some minor portion of the turbo-charger surge problem is related to the difficulties being encountered with the pitch scheduling port).on of the main engine control system. TDI should be required to assist and work closely with Mathers , Controls-to establish responsibility for and correct this situation.
- 6. Based on the above described performance, TDI should be put on notice that the de-rating work to date is unacceptable and payment withheld.
- Adequacy of the Engine De-Rating:
- 1. Based on a review of main engine historical maintenance and repair data and a comparison of engine component failure frequency and- mode with th'e modification accomplished as a result of the de-rating effort,
( it is anticipated that only minimal overall improvement in failure rates and time between failures cc overhauls f will occur. The most significant portion of this 1
v.
6-2
improvem2nt wi.11 occur for those components directly impacted by the improved combustion process which results from the increased availability of air blown for combustion.
- 2. It is believed that for the remainder of the engine component failures identified in Sections ~ '3.0 and 4.0 those not directly influenced by increased air flow, little or no change in failure rate, and probably no more than would be obtained by simply running the original engines at a reduced output with,out officially de-rating will occur. These component failures include:
- Cyliner heads -design and manufacturing defects
- Cylinder liner distortion and wear - due to block dis-tortion
- Piston ring distortion and wear -
due to block distortion .
- Cylinder blocks - distortion and cracking
- Connecting rod bearings - design of articulated connecting rod assembly
- Main bearings - premature wear, high 1:ading
- Cam shafts - premature wear i 3. It is estimated that when equated to dollars, the reduction in main engine maintenance and repair histor-ical average annuot cost resulting from de-rating 4
may approach twenty-five percent (25%).
6-3
- 4. Tha existing de-rated engines after incorporation of the additional modification identified in this report, can be kept running almost 'ndefinitely i if AMSH is willing to continue to maintain - them at the same comparatively high rate in terms of time and dollars.
- Additional Modifications:
Numerous additional modifications have been identified in Section 5.0 and should be incorpcrated to enhance the future reliable and efficient operation of 't h e de-rated engines. Some of the more important of these modifications are a result of, and not in addition
.to, the de-rating effort. The most significant of these is the turbocharger mismatch which should be rec tified by TDI by installing new matched turbochargers at no additional cost to the de-rating contract.
- Economic Evaluation of Re-engining of the M/V COLUMBIA:
- 1. Re-engining of the COLUMBIA for operation on Marine Diesel Oil, MDO, depending on the acquisition cost estimate / remaining vessel life combination considered, can offer a significant economic advantage over continued operation of the existing de-rated engines on MDO.
- 2. Re-engining of the vessel to operate on Heavy Tuel Oil, HF0, is a clearly superior economic alternative compared- to both re-engining for MD0 operation or 6-4
--~.- ...
I continued operation of the de-rated enginas on MDO, regardless o f- the acquisition cost / investment period combination considered in the economic analysis presented in Section 5.0.
Based. on the ' technical analysis and evaluation conducted and documented in this report and the results derived for the range of estimated re-engining acquisition cost / remaining vessel life combinations considered as - part of the economic analysis, it is recommended that the M/V COLUMBIA be re-engined for HFO operation at the earliest opportunity.
6-5
-,,-w .
- _ .-. w . . - . . - ~ , - p
c _ _ _ - . - - . - - _ - - _ - _ - - _ _
REFERENCES
'l . Delivery Order D.O. 707573 from the State of Alaska TDI for main engine derating _ new turbocharger installation and Trabon lubricating system in the .M/V COLUMBIA, 10/15/82.
- 2. M. Guralnick Associates, Inc., report, " Performance Predictions and Engine Selection Criteria for the M/V COLUMBIA, June 1982, S.O.A. Contract No. X61744.
- 3. TDI Report, Shipooard Test, M/V COLUMBIA, Starboard engine, S/N 72033, August 31, 1981.
- 4. TDI to Seaworthy Engine Systems, Inc.,
3/31/83, Typical C-17-123 compressor Transmittal performance dated map and Halter Marine DMRV-16-4 (DE C-17-173 turbos) test stand log sheet performance data.
- 5. Jarota, M.S. and Watson, M., Turbocharging the Internal Combustion Engine, 1982.
- 6. Jon O. Jacobson, Engine Rebuild Report - M/V COLUMBIA, March 31, 1981
- 7. G. Beshouri, J. Siegal, Inspection and Maintenance Report Port and Starboard Main Engines, M/V COLUMBIA, Trans-america DeLaval, Inc., November 1980.
- 8. Handbook for Selection of Marine Diesels, DeLaval Engine
& Compressor Division, 550 85th Avenue, Oakland, California 94621.
- 9. Stinson, K.W., " Diesel M.E., Engineering Handbook",
12th Edition, Business Journals, Inc., Stamford, Connecticut 1976.
- 10. Marine Diesel Standard Practices, Diesel Engine Manufac-turers Association, New York, New York, 1971.
- 11. Lichty, L.C.,
" Internal Combustion Engines", McGraw-Hill,
- 12. Modern Marine Engineers Manual, Vol. II, A. Osbourne, Editor, Cornell Maritime Press, Maryland, 1943.
- 13. Lamb, J., "The Rur.ning and Maintenance of the Marine Diesel Engine", 6th Edition, Charles Griffin & Company, Ltd., London, 1976.
- 14. Pounder, C.C., "!!arine Diesel Engines", Newnes-Butterworthys London, 1972.
- 15. Standard Handbook for Mechanical Engineers, T. Baumeister
& L.S . Marks, Editors, 7th Edition, 1967.
4
-Rdforences (continusd) l
- 16. .Zinner, K. , . " Supercharging of Internal ~ ~ Combustion = Engines'.', ,
' Springer-Verl'ag ,' Berlin , 1978.
- 17. . Northwest? Laboratories, " Damaged' Bearing-from Columbia",.
-Report No. E18022,. Seattle, Washington,.7-April 1980.
I e
s l
i-
. - - - . - . . .~ = . . . . - . .. . . .
CLOSSARY OF ENGINE RELATED TERMS AND COMMONLY USED FORMULAE 9
Piston Displacement - The cylinder volume in cubic inches swept by the pistons of an engine. It is equal to the number of cylinders times the area of each piston in square inches times the stroke in inches.
Piston Speed -
The total number of feet traveled by a piston in a given time interval, usually expressed in feed per minute.
It is sometimes called piston travel .
Horsepower (hp) -
A time rate of doing work. One U.S. (Knd British) horsepower is equal to 33,000 foot-pounds per minute.
One horsepower (metric) is equal to 75.0 kilogrameters per second. The relationship between U.S. and metric horsepower is:
One U.S. horsepower equals 1.014 metric horsepower One metric horsepower equals 0.9863 U.S. horsepower Indicated Horsepower (ihp) -
The horsepower developed in the cylinder. It can be determined from the mean indicated pressure, the engine speed and cylinder dimensions. The formula is shown in the Formula Appendix.
Mean Indicated Pressure (mip)- A defined, constant, hypothetical pressure which would deliver to the top of the piston in one stroke the same work as is actually delivered to the top of l the piston by the working fluid in one cycle. The formula is l shown in the Formula Appendix.
l e
_7w ,- , -- -r-- -----
Braka- Horsepower (bhp) - The horsepower delivered by the engine shaft at.'the . ou t pu t' end. The name is- derived from the fact that it was originally measured by a . brake d e vic e' . The formula is shown in the Formula Appendix.
Shaft Horsepower-- The net power available at the output ~ coupling of . 'a transmission system, such as propulsion' gearing ,' 'el'ectric propulsion system, . slip coupling, etc. It differs from the
. brake horsepower of the _ engine by the amount of losses in the-transmission device or system.
Brake Mean Effective Pressure (bmep) - A derived factor represented by' "P" when the PLAN formula is equated to BHP. It is also equal to the meand indicated pressure (MIP) multiplied by the mechanical efficiency expressed decimally. It cannot be measured directly. See'the Formula Appendix.
Torque -
A moment which tends .to produce rotation. It is the product of force and radius , . expressed. in, pound-feet or pound-inches. See the Formula Appendix.
Indicated Thermal Efficiency - The ratio of the heat equivalent of. one horsepower-hour to the number of heat units actually supplied per indicated horsepower-hour. This may be calculated
- from either the high or low heat value of the fuel, with proper designations as to which value is used. See the Formula Appendix.
1 Brake Thermal Effic ,acy a -
The ratic of the heat equivalent L of one horsepower-hour to the number of heat units actually l
l supplied per brake horsepower-hour. This may be calculated 1
. . . - . , . .~., - . . . . , -, , - , , - - , - . -
--v . , - - . , - + ~ . . - , -. , , , - . . - - . , , , , _ .
from either the high or low heat value of the fuel, with proper designation as to which value is used. See the Formula Appendix.
Mechanical Efficiency - The ratio of brake horsepower to indicated horsepower.
Turbocharger Surgin g -
Thephenomena arising during surging appear from the blower characteristic. Because of the - pulsating consumption of air, variations in prescure occur in the scavenging air receiver. The resulting pulsations act back through the air cooler and discharge pipe, which means that the impeller does not work against a uniform pressure, the result being that the amount of air from the blower will vary.
For example, if the air suction filter becomes contaminated, the amount of air through the blowers is reduced. This means that the operation point will move to the left on the ' blower characteristic, because the effect is the same as an increased resistance to the flow through the blower system and the curve for the flow resistance will then be higher. The result of the above-mentioned pulsations can be that the upper point on the blower characteristic will be e.ceeded, and the blower ceases to deliver air. The effect is that the flow resistance is reduced and the blower will again deliver air, and this alt'ernating effect will continue, i.e., the blower will not 1
work in a stable manner and will surge. !
Wi.h engines having more than one turbocharger delivering to the same scavenging air receiver, surging conditions will result in air being pressed backwards through the surging turbocharger
.._ .y..
by the rcmnining.turbochstgsrs.
The symptors of surging are:
l'. Unusual noise at the suction side' of the turbocharger
- can be'a. muffled but violent boom.
- 2. The amount of air sucked in by the turbocharger can vary a great deal -
can be confirmed by placing a piece of paper against the suction filter.
- 3. The pressure of the scavenging air in the receiver is considerably lower than normal and varies widely.
- 4. Sharp fluctuations in the air pressure drop during passage through the air filter.
Surging can often be prevented by lifting the safety valve on the scavenging air receiver, and at the same time, reducing the power of the main engine. The turbocharger system- must be cleaned at the earliest opportunity.
Surging can be caused by: ,
- 1. Contamination of elements in the turbocharger system.
- 2. Failure in the supply of energy to the turbocharger, for example, due .. o one or more of the engine cylinders not providing full power.
Bosch Smoke Number -
Bosch Smoke Number is an indication of the opacity (clarity) of the exhaust gases existing from a diesel engine as determined by the Bosch Smoke Meter. This is a filtering type smoke meter, usually portable, in which a primary sensor is used to collect a specific volume of exhaust gas by having it flow through a tab of filter paper. Any soot i
L-
c . .
-is trapped by the filter -paper. This- paper is than put into a photo-electric type reflection meter to determine the Bosch_ .'
Smoke Number. Generally, a reading of L.0 indicates a -slight
-hazing of the exhaust gas. Thus, readings falling wel l below l'0
. indicate a very clean, clear exhaust gas condition which is indicative of - more than adequate air flow into the engine.
l l . .
\ .
FORMULAE
- 1. -Horsepower per cylinder (any - reciprocating engine) i s:
4 hp = PxLxAxN 33,000 where P = Indicated mean effective pressure, psi; or brake
[ mean effective pressure, pst_(corresponds with thp or bhp)
L = Stroke of piston in feet f:
A = Net piston area sq. in.
, N = Number of power strokes per cylinder per minute 1
i 2. Brake Horsepower (test stand) is:
2 x n x r x rpm x W bhp 33,000 where r = Distance between the shaft center and the point of application of the weight to the brake arm, .
! in feet W = Effective weight on the brake arm in pounds rpm = Revolutions per minute of the brake shaft 1r = 3.1416 4
, 3. Horsepower per cylinder (any single-acting internal combustion engine) is:
hp = P x D x L x rpm 1
1 l
l
- - - - - - ,, - w--f. .-a.. e -,,an.- -, -.vw vv- '
t "v=* --rw'~' n~w-'~- '* - -~--s--~' ~~~~~-*~r
where hp = Horsepower per-cylinder.(bhp or thp)-
D P
= Mep in psi, Bmep or imep corresponds with bhp or ihp Diameter of cylinder bore in inches L = Length of stroke in inches
- C = 1,010,000 for four-cycle engines
- C = 505,000 for two-cycle engines 4 Brake'Mean Effective Pressure is:
bmep = bhp x 33,000 LxAxN -
where 4
bhp = Brake horsepower per cylinder and L, A and N are the same as mentioned in formula 1 for horsepower per cylinder.
- 5. Mean Indicated Effective Pressure is:
imep = ihp x 33,000 LxAxN where lhp = Indicated horsepower per cylinder and L, A and N are as mentioned in formula l for horsepower per cylinder.
- 6. Mean Effective Pressure is: '
p, hp x C
- A D x L x rpm where hp = Horsepower per cylinder (bhp or thp) i P
D
= Mep in psi, Bmep or imep corresponds with bhp or thp
= Diameter of cylinder bore in inches L = Length of stroke in inches
- C = 1,010,000 for four-cycle engines
- C = 505,000' for two-cycle engines Formulae (3) and (5) may be used for engines having cylinder
' ' dimensions in metric units, with modification of constants as follows: (The hp will still be in British or U.S. units of 33,000 ft-lbs per min.)
I
, - -,- _ - , - , - -,,.e -,. -w . - . , -
P = Psi as b: fore D = Diameter of cylinder bore in centimeters L = Length of stroke in centimeters
- C - 16,500,000 for four-cycle engines !
- C = 8,250,000 for two-cycle engines '
- 7. Brake Mean Effective Pressure is:
bmep = indicated mean effective X mechanical efficiency
, pressure (imep) ,
expressed declinally _
- 8. Torque in ft-lbs = Q = 5252 x hp where hp = Transmitted horsepower rpm = Rotational speed or shaft in revolutions per minute
- 9. Piston Speed = fpm = length of stroke in feet x rpm x 2
- 10. Indicated Thermal Efficiency - Et= g t
where, for oil Diesel engines H = High heat value of fuel used wt = Fuel consumption in Ib/ihp/hr or, for gas and dual fuel Diesel engin&s H = Heat value of fuel used (hhv for fuel and ihv for gas fuel) w t = Fuel consumption /ihp/hr (consumption in Ib for fuel oil and cu ft for gas) ll. Brake Thermal Efficiency = Eb" H 3
where, for dil Diesel engines H = High heat value of fuel used wb = Fuel consumption in lb/ bhp /hra or, for gas and dual fuel Diesel engines H = Heat value of fuel used (hhv for fuel and thv for gas fuel) wb = Fuel conspumption/ bhp /hr (consumption in Lb for fuel oil and cu ft for gas)
1
- 12. M chanical Efficiency in per cent =
fP x 100
- I
'13. Horsepower Requirements of Pumps:
(a) Circulating water pumps, for jacket water or raw ~
water systems, when total dynamic head is spect-fled in feet of water:
hp inp'ut = Spmxex H x where H = Total dynamic head expressed ln feet of water C = 3960 for fresh water (62.4 lb/cu ft)
C = 3855 for salt water (64 lb/cu ft) e = Pump efficiency, expressed decimally (b) Lubricating oil or fuel oil pumps:
hp input = apm t x P
where p = Discharge pressure, psi e = Pump efficiency, expressed decimally (with the discharge head expressed in psi, the constant 1720 is independent of variations in density of the liquid pumped. Horsepower capacities of oil pump mechanical drives or electric motors must be suffi-cient to start the pump with cold oil, usually assumed to have a maximum viscosity of 3000 SSU. The pump size must be selected to give the required capacity with hot oil, having a viscosity assumed to be 100 SSU.)
- 14. Specific fuel consumption correction factor for fuels of various high heat value:
! Factor = btu (hhv) 19,330
- where
- (hhv) = the high heat value of ;he fuel used I -
L
- Approx'imate values acceptable for computation. Note that the constant is based on fps system. Bore and stroke are given l
in inches and hp'is British or U.S. of 33,000 ft-lbs per min.
m --
-m m- w- -- - - ,,w-me
APPENDIX A
. TRIAL AGENDA M/V COLUMBIA, MARCH 24, 1983 e
SEA TRIAL AGENDA M/V COLUMBIA - MARCH 24,-1983 9
LLeave Pier 48 @ 07d0 HRS.
Enterprise Break-In:
Approx.
Eng. RPM Shaft RPM % Rated Pwr BHP Time 300 -
166.94- 40% 2500 1 HR Mid Power Cruise 330 183.63 53% 3300 1 HR Engine Performance Test:
360 200.33 69% 4300 , i HR Design Svc Cruise 385 214.24 85% 5248 1 HR MCR @ 100% 403 224.26 100% 6164 4 HR 10% Overload 403 224.26 110% 6791 1 tis Controls / Turbocharger Response Test:
(Mather Controls to supply)
Min. to include -
Bridge /ER/ Local Control Test /Tnansfer 1 HR (Slow) Ahead /Stop/ Astern 1 HR Full Ahead / Full Astern 1 HR 12 HRS Estimate Return Pier 48 - 1900 HRS.
l l
. l
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=
APPENDIX B M/V COLUMBIA MARCH 24-25, 1983 TRIAL DATA 4
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APPENDIX C M/V COLUMBIA SHAFT HORSEPOWER MEASUREMENT SEA TRIALS MARCH 24-25, 1983
_ _ . _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _. _.____.__._____)