Design and prototype of dual loop lubricant system to improve engine fuel economy, emissions, and oil drain interval

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Design and Prototype of Dual Loop Lubricant System To Improve Engine Fuel Economy, Emissions, and Oil Drain Interval

by

Michael J. Plumley B.S., Mechanical Engineering U.S. Coast Guard Academy, 1998

M.S., Mechanical Engineering Massachusetts Institute of Technology, 2005

MASSACHUSETTS INSTITUTE OQF TECHNOLOLGY

APR 15 2015

LIBRARIES

M.S., Naval Architecture and Marine Engineering Massachusetts Institute of Technology, 2005

Submitted to the Department of Mechanical Engineering in partial fulfillment of the requirements for the degree of

DOCTOR OF PHILOSOPHY IN MECHANICAL ENGINEERING

AT THE

MASSACHUSETTS INSTITUTE OF TECHNOLOGY

February 2015

Signature redacted

Signature of Author:

Signature redacted

6D'epartmdit of Mechanical Engineering January 9, 2015

Certified by:

Accepted by:

Wai K. Cheng Professor of Mechanical Engineering Committee Chair

_Signature redacted______

Victor W. Wong Prinpal Research Scientist-thd Lecturer of Mechanical Engineering

Signature redacted

Thesis Supervisor

David Hardt Professor of Mechanical Engineering Chairman, Committee for Graduate Students

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Design and Prototype of Dual Loop Lubricant System To Improve Engine Fuel Economy, Emissions, and Oil Drain Interval

by

Michael J. Plumley

Submitted to the Department of Mechanical Engineering on January 9, 2015 in Partial Fulfillment of the Requirements for the

Degree of Doctor of Philosophy in Mechanical Engineering

ABSTRACT

Regulations aimed at improving fuel economy and reducing harmful emissions from internal combustion engines place constraints on lubricant formulations necessary for controlling wear and reducing friction. Viscosity reduction results in fuel economy improvement, with benefits of up to three percent reported in some studies. Such reductions are limited by engine durability constraints. Recent limits on oil additives, driven by emissions aftertreatment requirements, impose additional design tradeoffs.

The benefit of segregating lubrication systems, in light of modern formulation constraints, is investigated through modeling and experiment. Many findings are applicable to spark and compression ignition engines, with an emphasis placed on diesel engines, given the

implementation of the first heavy duty diesel fuel economy regulations. Nearly all engines used today employ a lubrication system with a pump delivering an oil to all engine regions. Axiomatic design concepts are applied to describe the associated design tradeoffs. Two dual loop prototypes were developed, incorporating independent oil systems for the engine valve train and power cylinder, decoupling many lubricant functional requirements. Oil analysis and friction

measurement were used to quantify performance. A combination of high viscosity lubricant in

the valve train, with low viscosity in the power cylinder, increased fuel economy while maintaining wear protection. Effective protection of subsystems from contamination and oil degradation, particularly the elimination of soot in the valve train, was demonstrated.

Detailed friction and oil composition modeling was used to investigate opportunities for further friction and wear reduction. Techniques for investigating oil composition changes along the liner in modern friction models are developed. Differences in lubricant functional requirements along the liner are highlighted. Model results indicate that vaporization along the liner increases lubricant viscosity near piston top dead center, providing a potential wear reduction benefit.

Thesis Supervisor: Victor W. Wong

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ACKNOWLEDGEMENTS

I would like to thank several people for making the success of this project possible. First, my

thesis advisor, Dr Victor Wong, for his guidance and freedom to explore the topics I found most interesting. I also appreciate the advice and patience of my thesis committee, Professor Wai Cheng and Professor Bill Green, and the many faculty and staff members at MIT who provided insight and guidance along the way.

I am indebted to the US Coast Guard for funding and support for this incredible opportunity, and

specifically my colleagues at the Coast Guard Academy who provided continued encouragement and technical advice. This study was part of a larger study at MIT for the Department of

Energy's Vehicle Technologies (VT) Program in the Advanced Fuels and Lubricants group, entitled "Lubricant Formulations to Enhance Engine Efficiency in Modern Internal Combustion Engines". I greatly appreciate the hard work, inspiration, and talent of my co-reseachers and friends Tomas Martins and Mark Molewyk. This study also benefited from discussions and inspiration of Dr Soo Youl Park, Grace Gu, and Ian Tracy. We enjoyed the support of various industry partners, including Bob Lindorfer of Kohler Engines, Luigi Arnone of Lombardini, Dr Jai Bansal, Maryann Devine, and Dr David Brass of Infineum, and David Atherton of Detroit Diesel.

I'd like to thank all members of the Sloan Lab, without which this work would not be possible.

Thane Dewitt and Raymond Phan for their continued help in the shop, Janet Maslow for the exceptional support, and Alex Sappok for his advice. I'd like to acknowledge my lab mates' advice and comradre, including Kevin Cedrone, Justin Ketterer, Felipe Rodriguez, Eric Senzer, James Jorgenson, Tim Murray, Nick Custer, Paul Folino, and Eric Zhangi. Thank you to my Course II colleagues, including my study partners and friends: Leo Banchik, Subarna Basnet, Ashley Morishige, Ashwin Raghavan, Matthieu Picard, Stephanie Nam, Yi Liu, Sasha Miao, and Mark Jeunnette. Thanks also to the Course II community, including Leslie Regan and Joan Kravit. This list is not exhaustive, and I wish the best of luck to all I had the opportunity to meet and work with at MIT.

Most of all, to my family and past teachers and mentors, wherever I had the honor to get to know you, in schools, offices, at sea and ashore. Thank you for your dedication and support over all these years. Finally, a special thanks to Malissa, for her unwavering support and inspiration.

Michael J. Plumley January 2015

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LIST OF FIGURES

Figure 1 Adjusted light vehicle fuel economy historical trends, in miles per gallon (MPG), and

projected values based on new NHTSA rules... 24

Figure 2 Diesel engine contributions to harmful emissions... 26

Figure 3 United States diesel fuel sulfur limits 2005-2014. ... 27

Figure 4 US heavy duty diesel particulate matter emission limits (data from [21])... 27

Figure 5 US heavy duty diesel allowable NOx limits (data from [21])... 28

Figure 6 Conventional (left) and dual loop (right) lubrication system configurations... 31

Figure 7 Lubricant system for 3 cylinder Kohler KDW 1003 (turbocharger removed) [30] [31]. A similar configuration, the KDW 702, a twin cylinder naturally aspirated diesel engine, was u sed in th is stud y ... 3 2 Figure 8 Impact of OCP viscosity modifier concentration on base oil viscosity [56]... 35

Figure 9 Comparison of viscosity temperature relationship for three different lubricants form ulated from different base oil groups. ... 36

Figure 10 Axiomatic design matrix describing conventional engine lubricating system... 39

Figure 11 Axiomatic design matrix for dual loop lubricating system. ... 41

Figure 12 Cylinder head bench test rig shown without insulation or shaft guards... 42

Figure 1-3 Fore and aft views of engine test cell rig. Labeled photos are available in Appendix C. ... 4 3 Figure 14 Dual loop lubrication system for test cell engine in split configuration... 45

Figure 15 Front panel, main tab, for test cell engine LabVIEW user interface... 46

Figure 16 Effect of crank angle measurement error on NIMEP estimates... 50

Figure 17 Detailed NIMEP vs. FMEP data with uncertainty bars for precision error... 51 Figure 18 Comparison of NIMEP and FMEP for power cylinder lubricant additive parametric

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Figure 19 Changes in FMEP over time for same lubricant. ... 54

Figure 20 Day to day variation in NIMEP vs. FMEP for fired test cell engine... 55

Figure 21 Soot sampling system installed on test cell engine. ... 56

Figure 22 Top down view of light tower with modified KDW 702 installed... 57

Figure 23 Cylinder head with modified drain lines on either side... 58

Figure 24 Oil drain installed just above crankcase vent. On the right, crankcase vent is routed through a separator prior to the intake. ... 59

Figure 25 Custom aluminum valve train oil sump... 60

Figure 26 Dual loop lubrication system for light tower engine... 61

Figure 27 Mechanical valve train oil pump mounted on camshaft power take off (PTO). ... 61

Figure 28 The mobile aging unit was operated continuously for 10-20 hour periods with short breaks for refueling. Two lights were energized to maintain steady state operation... 62

Figure 29 Viscosity Temperature relationship for lubricants used in study... 64

Figure 30 Shear rate vs Viscosity for Newtonian 40 and 15W40 lubricants at 1500C... 65

Figure 31 Stribeck curve depicting the relationship between load (W), speed (U), and viscosity (p) on the friction coefficient between sliding parts [56]... 67

Figure 32 General breakdown of engine energy distribution (left) [741 and component m echanical losses (right) [151... 68

Figure 33 Comparison of density for a lubricant used in this study with Dowson Higginson re latio n . ... 6 9 Figure 34 Film thickness within 30 degrees of TDC under top ring. Model results for the MM I I engine model with 15W40 lubricant at A 100 operating condition... 76

Figure 35 Temperature distribution under top ring for two engine operating conditions [56]... 78

Figure 36 Assumed torque speed curve for the MM1 I model, based on a modern 11.0 L turbocharged diesel engine. ... 79 Figure 37 Comparison of temperatures under power cylinder components over the compression

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Figure 38 Minimum film thickness, hydrodynamic, and boundary friction forces under top ring for A 100 operating condition (left). On the right, the related instantaneous power losses [5 6 ] ... 8 1

Figure 39 Hydrodynamic and boundary losses under top ring during A100 condition [56]... 82 Figure 40 Hydrodynamic and boundary friction losses under top ring for A 100 condition (left)

and associated peak wear factor (right) [56]... 83

Figure 41 Boundary and hydrodynamic power losses vs. position for A100 condition with

15W 40 and 5W 20 [56]... 83

Figure 42 Distance along liner experiencing boundary friction for various multigrade lubricants [5 6 ] ... 8 4 Figure 43 Power losses for top ring and skirt with respect to position along liner [56]... 85 Figure 44 Power losses under the top ring and skirt with respect to temperature [56]... 85 Figure 45 Viscosity vs. temperature profile with designated fuel economy and wear reduction

re g io n s [5 6 ]... 8 6

Figure 46 Viscosity temperature profiles with varied wear reduction index [56]... 87 Figure 47 Friction and wear effect from changing wear reduction and fuel economy indices. ... 88

Figure 48 Top ring power losses for the 15W40 and VM4 viscosity profile cases [56]... 89 Figure 49 Top ring hydrodynamic and boundary friction losses resulting from use of 15W40 or

5W20 with TBC. Taken from [123] with permission... 90

Figure 50 Effect of shear thinning over one cycle for a I 0W30 lubricant assuming parameters for the K D W 702 at rated speed... 92

Figure 51 Effect of shear thinning on average friction power loss for a 10W30 lubricant

assuming parameters for the KDW 702 at rated speed... 93

Figure 52 Total engine friction FMEP, with exception of PMEP. 15W40 in valve train and

various base oils in pow er cylinder... 94

Figure 53 Valve train friction values for cases given in last figure. Uncertainty based on cycle to cycle variation for each 50 cycle testpoint shown. ... 95

Figure 54 Power cylinder NIMEP vs. FMEP results at 2400 rpm... 96

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Figure 56 Brake specific fuel consumption with various lubricants... 99 Figure 57 Cylinder head arrangement, KDW 702 engine used in experimental work in this study.

C am shaft show n for reference. ... 105

Figure 58 Overhead cam configuration for KDW 702engine used in experimental work in this stu d y ... 10 6

Figure 59 Rocker arm friction over a single engine cycle for KDW 702 model with 15W40 lu b rican t at 80 'C . ... 10 9

Figure 60 Maximum Hertzian pressure under cam follower over one cycle for different engine speeds for K D W 702 m odel... 1 10

Figure 61 Valve train losses vs. speed for different multigrade lubricants. ... Ill

Figure 62 Model effects of shear thinning on valve train losses. ... I I

Figure 63 Valve train torque as a function of speed with and without fuel injectors... 112

Figure 64 Effect of temperature and valve train branch on valve train friction. Bench test and m o d el results sh o w n ... 113

Figure 65 Valve train friction vs. temperature, 10W30 lubricant, 1200 camshaft rpm... 114

Figure 66 Valve train losses with and without fuel system on test cell engine. ... 115

Figure 67 Valve train losses vs. speed for different multigrade lubricants, cylinder head bench te st... 1 16

Figure 68 Fired valve train friction results for various lubricant combinations in test cell engine. E ngine speed 1800 rpm ... 117

Figure 69 Valve train friction benefit from Newtonian 40 vs. 15W40... 118

Figure 70 Effect of oil supply pressure on camshaft torque. ... 119

Figure 71 Relationship between valve train oil supply pressure and resulting pump output power.

... 1 2 1 Figure 72 Compare FMEP for common lubricant in each subsystem at 1800 rpm, 3.5 bar

co n d itio n . ... 12 2 Figure 73 Common lubricant in each subsystem, Group II/II1 multigrade and Group IV

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Figure 74 NIMEP vs. total engine FMEP for 1800 rpm tests... 124

Figure 75 Compare FMEP for various lubricant combinations in the head/power cylinder. Data labels provide difference in FMEP from the 15W40/15W40 baseline case... 125 Figure 76 Dual loop friction results, fired test engine 5W20 to NEW40 lubricants at 2400 engine

rpm and N IM E P 5.0 bar... 127

Figure 77 Prototype dual lubricating loop engine installed in commercial light tower... 135

Figure 78 Kinematic viscosity at 1000C for valve train and power cylinder lubricants. Error bars

are based on repeatability values reported in ASTM D445 and ASTM D6616... 140

Figure 79 Total base number and total acid number for valve train and power cylinder lubricants.

... 14 1 Figure 80 Additive concentrations for valve train lubricant. Change at 250 hours shown in

legend. Error bars for P and Ca given based on ASTM D5185 repeatability... 142 Figure 81 Additive concentrations for power cylinder lubricant... 142

Figure 82 TEOST-MHT results for valve train and power cylinder lubricants, as compared to new lubricant, after 250 hours. ... 143 Figure 83 FTIR spectrum for valve train lubricant. Figure provided by Savant Laboratories... 144

Figure 84 FTIR spectrum for power cylinder lubricant. Figure provided by Savant Laboratories.

... 14 4 Figure 85 Oil samples at 50 and 250 hours. New oil is on the left of each photo. The valve train

sample is in the center, and power cylinder lubricant sample on the right... 146

Figure 86 Valve train wear and contaminant concentration over time... 147 Figure 87 Power cylinder wear and contaminant concentration over time. ... 147

Figure 88 Additive concentrations for valve train lubricant measured at 250 hours compared to a new oil doped with fuel. New oil with 5% fuel concentration is also shown... 150

Figure 89 Raw particulate matter emissions from 2002 Cummins ISB using a low and high sulfur lubricant. D ata from [16]. ... 154

Figure 90 Ash emission for different oil consumption rates and oil chemistries from 2002 Cummins ISB. Data from [16]. Lube oil consumption for the test conditions were 5.8 and

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Figure 91 Typical layout of modern diesel engine with EGR and aftertreatment system utlizing

D O C , D P F , and S C R . ... 157

Figure 92 Elemental composition of soot samples taken under 2400 rpm operating condition u sin g IC P -A E S ... 1 6 2 Figure 93 Soot emission with various lubricants in power cylinder of test cell engine. ... 163

Figure 94 O il thickness near top ring TD C [56]... 165

Figure 95 Average molecular weight along liner for assumed 15W40 lubricant composition [56]. ... 1 6 6 Figure 96 Viscosity along liner with and without vaporization effects [56]. ... 167

Figure 97 Power loss with and without vaporization correction. ... 168

Figure 98 Temperature effect on composition molecular weight near TDC... 169

Figure 99 Friction Vaporization model flow chart. ... 169

Figure 100 Extrapolated viscosity temperature relationships for higher molecular weight base o ils ... 17 0 Figure 101 Viscosity temperature comparison between measured data (dark blue) and model extrapolation (blue star) for a C 33 PA O ... 171

Figure 102 Composition, by carbon number, and predicted viscosity for assumed composition. Carbon number, mass fraction, and estimated boiling and melting points of 10 species listed along top of plot on right. Estimated A and B coefficients for Walthers Equation also given. Solid line show s a 15W 40 for com parison. ... 172

Figure 103 Proposed base oil formulation with light and heavy species. Assumed viscosity temperature profile of each species shown on right... 173

Figure 104 Mass composition and viscosity temperature profile of proposed formulation follow ing vaporization depletion at TD C . ... 173

Figure 105 Viscosity temperature profiles along liner when accounting for vaporization... 174

Figure 106 Viscosity temperature profile for a viscosity modifier product (green) and a Group II b ase (b lu e )... 17 5 Figure 107 Specific viscosity as a function of viscosity modifier concentration. ... 176

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Figure 108 Change in specific viscosity for different mass concentration of OCP in a subject gro up II b ase o il... 17 6

Figure 109 Fuel dilution effects were analyzed along the liner assuming blending of fuel and oil

on a m ass basis along the upper 25% of the liner... 177

Figure 110 Effect of fuel dilution on upper quarter of liner on viscosity over one engine cycle. M ass fraction of fuel denoted on the right... 178

Figure 111 Top ring friction power loss with fuel dilution on upper quarter of liner. Model parameters are those for the KDW702 at rated speed... 178

Figure 112 Average film thickness change on upper 25% of liner due to fuel dilution... 179

Figure 113 GT SUITE valve train model schematic. Camshaft shown for reference... 194

Figure 14 Model valve train branch component configuration in GT SUITE. Number 2 cylinder in take valv e show n ... 194

Figure 115 Detailed data. NIMEP vs. total engine FMEP for 2400 rpm tests. ... 196

Figure 116 Detailed data. NIMEP vs. total engine FMEP for 1800 rpm tests. ... 197

Figure 117 Test cell engine, forw ard looking aft... 198

Figure 118 Test cell engine, aft view looking forward... 198

Figure 119 Test cell engine looking forw ard... 199

Figure 120 An early protoype for the mobile aging right with electric valve train oil pump. Valve train oil sam ple valve not installed on this prototype. ... 199

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LIST OF TABLES

Table 1 A PI Base Stock Group Definitions ... 33

Table 2 Engine operating conditions at steady state... 49

Table 3 Lubricant specifications for base oil comparison. ... 63

Table 4 Lubricant specifications for base oil comparison with different shear rates. ... 65

Table 5 Lubricant specifications for power cylinder additive effect comparison trials. ... 66

Table 6 Summary of component contributions to piston cylinder friction in literature ... 74

Table 7 MMI I engine parameters used in modeling study [56]. ... 78

Table 8 Engine parameters, measured and assumed, for KDW702, used in modeling [56]... 92

Table 9 Valve train power losses assuming different computation modes... 107

Table 10 Valve train component power losses (fuel system not included). ... 108

Table 11 Valve train branch component power losses. ... 109

Table 12 FMEP results for 1800 rpm condition with relative difference for common vs. different lu b ric a n ts... 12 6 Table 13 Test condition range for 250 hour oil aging test with mobile aging rig ... 136

Table 14 Viscosity data from 250 hour dual lubrication loop field test including fuel dilution. 139 Table 15 FTIR results for valve train and power cylinder field test... 145

Table 16 Thermogravimetric results by ASTM El 131 ... 146

Table 17 W ater by Karl Fischer, ASTM D6304A ... 148

T able 18 Fuel doped sam ple results... 15 1 Table 19 Detailed model parameters for MMI I and KDW 702 top ring friction models. ... 195 Table 20 FMEP results for 2400 rpm condition with relative difference for common vs. different

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Table 21 FMEP results for 1800 rpm condition with relative difference for common vs. different lu b rican ts... 1 9 7

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NOMENCLATURE

API American Petroleum Institute

BDC Bottom Dead Center

BHP Brake horsepower

BMEP Brake Mean Effective Pressure

BSFC Brake specific fuel consumption

Ca Calcium

CI Compression Ignition

CO Carbon Monoxide

CO, Carbon Dioxide

Cu Copper

Deg, Degrees

DOC Diesel Oxidation Catalyst

DP Design Parameter

DV, p Dynamic Viscosity DPF Diesel Particulate Filter

ECM Engine Control Module EGR Exhaust Gas Recirculation

EHD Elastrohydrodynamic

FE Fuel Economy

Fe Iron

FM Friction Modifier

FMEP Friction Mean Effective Pressure

FR Functional Requirement

GIMEP Gross Indicated Mean Effective Pressure

HC Hydrocarbons

HTHS High Temperature High Shear

ICP-AES Inductively Coupled Plasma - Atomic Emissions Spectrometry KV, v Kinematic Viscosity (KV40 = Kinematic Viscosity at 40'C)

LNT Lean NOx Trap

MEP Mean Effective Pressure

Mg Magnesium

MoDTC Molybdenum Dialkyldithiocarbamate

MPG Miles Per Gallon

NHTSA National Highway Traffic Safety Administration

NIMEP Net Indicated Mean Effective Pressure

NOx Oxides of Nitrogen

NPT National Pipe Thread

OCP Olefin Copolymer

OEM Original Equipment Manufacturer

P Phosphorous

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Pb Lead

PAH Polyaromatic Hydrocarbons

PAO Polyalphaolefin

PC Power Cylinder

PM Particulate Matter

PMA Polymethacrylate

PPB Parts per billion

PPM Parts per million

PSI Pounds per square inch

Re Reynolds Number

RPM Revolution Per Minute

S Sulfur

SAE Society of Automotive Engineers SCR Selective Catalytic Reduction

SI Spark Ignition

SLPM Standard Liters per Minute

SUV Sport Utility Vehicle

S02 Sulfur Dioxide

SOF Soluble Organic Fraction

SOHC Single Overhead Cam

SSI Shear Stability Index

STLE Society of Tribologists and Lubrication Engineers

TAN Total Acid Number

TBC Thermal Barrier Coating

TBN Total Base Number

TCP Tricresyl Phosphate

TDC Top Dead Center

TEOST-MHT Thermo-Oxidation Engine Simulation Test - Moderately High Temperature

TGA Thermo-Gravimetric Analysis

ULSD Ultra Low Sulfur Diesel VI Viscosity Index Improver

VM Viscosity Modifier

VOF Volatile Organic Fraction

VT Valve Train wt Weight XRD X-Ray Diffraction ZDDP Zinc Dialkyl-Dithio-Phosphate Zn Zinc rum y Mechanical Efficiency Shear Rate Micrometer 22

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Chapter 1 INTRODUCTION

1.1 Motivation

This study was motivated by the need for improved fuel economy and reduced harmful

emissions from diesel internal combustion engines. Reducing wear and oil change intervals are related motivations. Options for improving fuel efficiency and emissions are many, and already fill volumes. This study focuses specifically on the engine lubrication system. Common to all

internal combustion engines, the lubrication system serves two primary purposes, reducing friction and controlling engine wear and corrosion [1]. Friction reduction has a direct benefit on fuel economy, as mechanical efficiency is on the order of only 90% for today's engines.

Migration of oil to the exhaust stream due to oil consumption has a significant impact on emissions, allowing for gains through proper formulation.

This study was part of a larger study at MIT for the Department of Energy (DOE) Vehicle Technologies (VT) Program in the Advanced Fuels and Lubricants group, entitled "Lubricant Formulations to Enhance Engine Efficiency in Modern Internal Combustion Engines". Much of

the work is summarized in the VT Program annual reports [2][3]. That study was aimed at optimizing lubricant formulations to improve mechanical efficiency and emissions. Specifically

by: 1) identifying optimal lubricant formulations for particular engine subsystems through

friction modeling and experiment; 2) improving formulation through friction modeling and experiment incorporating local oil composition models; 3) quantifying the potential emissions and efficiency benefits of operating engines with independent cylinder head and crankcase lubricating systems. As part of that study, the dual loop lubricating system prototype was developed to study the effect of lubricant formulations in particular engine subsystems. This work details those benefits, as well as others that may be gained by incorporating such an engine design.

1.1.1 Fuel Economy

Improved efficiency has long been the goal of engine design. The most obvious manifestation of this is in worldwide fuel efficiency standards for passenger cars and heavy duty vehicles. Recent decisions in the US are aimed at significant improvements in fuel economy, motivated by reducing the country's dependence on foreign oil sources. In August 2012 the U.S. National Highway Traffic Safety Administration (NHTSA) issued its latest ruling covering standards to reduce Corporate Average Fuel Economy (CAFE) targets (for model years 2017-2025) under the Energy Policy and Conservation Act [4]. Rules for model years 2012-2016 were issued in 2010. Estimates are the 2012 rule will result in a near doubling of fuel efficiency of U.S. passenger cars and light vehicles by 2025, to 54.5 mpg on average [5].CAFE standards were first enacted in

1975 to regulate passenger car fuel economy. Light vehicles represent passenger cars, sport

utility vehicles, minivans, and all but the largest pickup trucks and vans. Adjusted values for miles per gallons are shown in Figure 1. The figure values are taken from [6], and include values

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adjusted for real world driving conditions. These conditions, evaluated by the United States Environmental Protection Agency (EPA), account for various factors including speed, aggressive driving habits, air conditioning use, and cold temperature operation. Levels remained relatively flat, even falling, from 1980 to 2005, at which time regulations began to change. The data is affected by EPA estimates of real world driving conditions, with those definitions evolving over time and not retroactively accounted for all years [6]. The change in classification between passenger vehicles, light trucks, and SUV's is propagated back. Figure 1 also shows projected fleet wide values anticipated in [7].

US Light Vehicle Fuel Economy - Adjusted Values

60 Projected - 2012 Rule 50 40 --- --- ~ -- -- U ---- _ 30 Historical Data 20 1 10 1970 1980 1990 2000 2010 2020 2030

Figure 1 Adjusted light vehicle fuel economy historical trends, in miles per gallon (MPG), and projected values based on new NHTSA rules.

Diesel fuel represents a major industry cost and has recently become the target of fuel economy regulation. It is estimated to account for 40% of the total operating cost of Class 8 trucks in North America [8] and 12% of all US oil consumption [9]. Other estimates place diesel fuel costs at 25-35% of total haulage firm overheads [10]. Beyond company cost, heavy duty truck

transport is a significant component of US oil consumption and economic interest. 87% of freight hauled, by value, is by heavy duty diesel, with similar levels anticipated as far into the future as 2035 [11]. Heavy duty trucks are the fastest growing sector for US greenhouse gas emission, and currently account for 22% of overall emissions [11]. In 2011 the first ever U.S. medium and heavy duty vehicle fuel efficiency standards were also announced, affecting

2014-2018 model years, with a goal of efficiency improvements for combination tractors ("big rigs")

of 20%, and up to 15% for heavy duty pickups [12][13][9]. As of 2015 the EPA is developing requirements for Phase 2 of the heavy duty diesel fuel consumption reduction. Under the 2011

rule heavy duty pickups have a mile per gallon and CO2requirement, where combination tractors

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research under the Supertruck program resulted in a 75% increase in class 8 fuel truck economy through a variety of means [14].

Fuel economy changes have prompted the first direct impact on lubricant formulation. The focus on fuel economy in 2011 prompted the Engine Manufacturer's Association to request the

American Petroleum Institute's (API) Diesel Engine Oil Advisory Panel to develop a new category, which is currently designated PC- 11 [8]. PC-l will contain two categories, one of which will be backward compatible, and the other a "fuel economy" formulation. The fuel economy will be achieved by setting the first viscosity limits on HTHS. While it has yet to be identified, industry expectations are that PC-Il will result in an HTHS 150 limit of 2.9-3.2 cP for the fuel economy specification [10]. A backward compatible lubricant, with a higher HTHS of

3.5 cP is also expected. Under SAE J300 HTHS 150 minimum limits are currently 3.7 cP for

15W40 lubricants, with typical lubricants exhibiting HTHS 150 values over 4.0 cP.

Engine friction accounts for a significant portion of fuel energy losses in on road vehicles. Richardson estimated mechanical friction accounted for 4-15% of total energy losses [15]. Recent estimates confirm that mechanical losses are 13-16% of total losses, with half of these associated with mechanical friction [10]. Reports from Infineum indicate fuel economy gains of up to 2% are possible in "line haul" mode as a result of the 33% friction reduction feasible from lubricant selection. At low speed, low load, these benefits increase to 3%, with even greater benefits at idling conditions [10]. As a result, less efficient operating conditions lead to greater opportunities for fuel economy improvement for an optimal formulation.

Mechanical efficiency is a measure of that portion of the gross indicated power used to do useful work. It is the ratio of the brake power to the indicated power as defined in [1] and shown in Equation (1), where Pb is the brake power and Pig is the gross indicated power. It may be

expressed in terms of mean effective pressure (MEP), which allows comparison of engine parameters independent of their displacement. The indicated mean effective pressure, IMEP, is the average pressure in the combustion chamber over a cycle. If considering all four strokes, the indicated work is defined as the NIMEP, or net indicated mean effective pressure. If exhaust and intake strokes are neglected, the value given is the GIMEP, or gross mean effective pressure, which represents the conversion of combustion energy to work without the losses from these strokes. The difference between NIMEP and GMEP is the PMEP, or pumping mean effective pressure. The brake mean effective pressure is designated as BMEP. The definition of friction mean effective pressure, FMEPg, is given as the difference between GIMEP and BMEP.

Pb _-q _{1 -} FMEP

-(1) m Pif GIMEP

In this work, the primary interest is in those friction components effected by lubricants, so FMEP, as presented in the following chapters, is given as the difference in the NIMEP and BMEP, eliminating the effect of pumping losses, which, although significant, are not expected to be affected by engine lubricants.

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1.2 Emissions

Any modem study of lubricant optimization must also include consideration of emissions

impacts. Lubricants which improve fuel efficiency can have a negative impact on emissions. For heavy duty on road diesels regulated emissions in the US include particulate matter, NOx, and hydrocarbons [16]. Emissions regulations have influenced lubricant development as a result of fuel sulfur reduction, motivated by reductions in sulfur dioxide, and additive limits, motivated by particular matter and nitrogen oxide (NO,) reduction. Diesel engines consume fuel and, to a far lesser extent, lubricating oil, which contribute to engine out emissions of particulate matter (PM),

SO2, NO,, HC, CO, and CO2. Specific power levels for diesel engines in the on road fleet have

increased significantly over the last 40 years; however recent emissions regulations have slowed that increase. Regulations of heavy duty diesel engines began with NOx, PM, HC, and CO limits on a 13 mode steady state test in 1978 [17].

air

Engine

DPF

Engine-Out Clean oilEmissions Emissions fuel PM, SO2, NOX HC, CO, C02

Figure 2 Diesel engine contributions to harmful emissions.

Sulfur -The US began regulating diesel fuel sulfur content in the 1990s. In 1994 highway fuel sulfur was limited to 500 ppm, or low sulfur diesel (LSD), followed by a reduction to 15 ppm, or ultra-low sulfur diesel (ULSD), in 2006. Other fleets followed in later years, with all diesel fuels reduced to ULSD by 2014 as indicated in Figure 3 [18]. Sulfur in diesel fuel was attributed to increased SO, and particulate matter emissions [161. SO,, along with NO,, is a contributor to acid rain. NO2 formation from NO occurs in the atmosphere, exhaust, and cylinder in various

degrees. NO2 can react with H20 in the upper atmosphere to form Nitric acid, HNO3. Likewise SO2 formed in cylinder combustion reacts with local 02 to form SO3 and then with H20 to from

Sulfuric acid, H2S04.In addition to the harmful effects of these emissions in the atmosphere, sulfur compounds have the added drawback of poisoning aftertreatment catalysts. Sulfur in diesel lubricants may also be shown to increase particulate matter emissions as well as diesel ash emissions, which lead to increased DPF maintenance requirements [16]. In exhaust chemistries with high Sulfur content the condensing of Sulfuric acid can lead to the formation of particulate matter separate from that adsorbed directly on Carbon. Sulfur is a leading cause of increased PM levels in high Sulfur versus low Sulfur fuels. Past studies identified reductions in fuel Sulfur correlate well to reductions in PM [16].

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US Diesel Fuel Sulfur Limits (ppm)

2005 2006 2007 2008 2009 2010 2011 2012 2013 2014

Highway 0

Non-road Soo sm Soo

Marine/ Locomotive Soo 5W Soo Soo

Non-road Marine/Locomotive

Figure 3 United States diesel fuel sulfur limits 2005-2014.

Particulate Matter -The presence of polyarornatic hydrocarbons (PAH) and other volatile organic compounds have led to the identification of PM as a known carcinogenic. Particulate of 10

micron and smaller, PM10, is particularly dangerous to health due to its effects on the lungs. As a

result regulations have led to a reduction in particulate matter over the years. In 2007 particulate matter limits were reduced by 90%. Allowable particulates in diesel fuel have been significantly reduced by regulation since the 1980's. Limits for heavy duty diesel vehicles, characterized as those with a gross vehicle weight rating of over 8500 lbs, are given in Figure 4. In 2007

particulate matter was reduced by 90%, resulting in a need for control strategies beyond in engine techniques. As a result, diesel particulate filters (DPF), which capture up to 99% of

particulate emissions, have been installed on nearly all new heavy duty diesel engines in the US and Europe since 2007 [19]. Maintenance intervals and performance of these filters are affected

by lubricant oils. Ash residue can lead to the need for in shop cleaning after certain periods [20].

Specifically, metallic components in lubricants contribute the majority of ash found in DPF, with sulfur in lubricants and fuels contributed to increased levels of ash due to sulfation effects L16].

US Heavy Duty Diesel Allowable Particulate Matter

0.7 0.60 0.60 0.6 0.5 1 .0.4 0.3 ho. 0.25 0.2 0.0 0.10

0.1 E m

0.01 0.01 0 "" 1"" 1985 1988 1990 1991 1994 1998 2007 2010

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NO, -NO, is the primary contributor to photochemical smog and Ozone through reaction with

sunlight, unburned hydrocarbons, and atmospheric Oxygen. NO, was the last of major criteria pollutants to receive significant regulatory attention and the latest of the heavy duty diesel emissions to receive a significant regulatory limit as shown in Figure 5. Along with SO,, it is also a contributor to acid rain. As with particulate matter reductions in 2007, the 2010 change was of an order of magnitude, requiring significant aftertreatment. The use of advanced aftertreatment technologies, such as diesel oxidation catalysts (DOC), selective catalytic

reduction (SCR), and lean NO, traps (LNT), were required to achieve the new limits in the US as well as Europe and Japan.

US Heavy Duty Diesel Allowable NOX

12 10.7 10.7 10 8 6.0 6.0 . 6 - - - --- 5.0 4.0 4.0 0.2

0

₁₉₈₅ ₁₉₈₈ ₁₉₉₀ ₁₉₉₁

IlL

₁₉₉₄ ₁₉₉₈ ₂₀₀₇ ₂₀₁₀

Figure 5 US heavy duty diesel allowable NOx limits (data from [211).

Phosphorus and sulfur based compounds used in some friction modifiers can poison emission aftertreatment systems [16]. This has additional implications for PC-Il formulation, as increased restrictions for gasoline engines require phosphorous content below 0.08 %,, to claim universal oil application to both diesel and gasoline fleets. In the past, formulations of 0.12%,, were allowed [8].

Non-road engines - In 2014 Tier IV regulations take effect for non-road engines [22]. These standards effectively set limits for non-road engines similar to those for the on road fleet, bringing about an order of magnitude reduction in PM and NOx emissions from the tier 3 standards. For 130 kW engines PM and NOx emissions are reduced from 0.2 to 0.02 g/kWh and 4.0 to 0.4 g/kWh respectively. These limits effectively require the use of aftertreatment systems on non-road engines [23].

Result - As of 2014 diesel particular filters, diesel oxidation catalysts (DOC), exhaust gas recirculation, and SCR with urea are employed on most on and off road diesel engines in the US,

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Japan, and Europe. In addition, diesel particulate filter use is now employed on marine and diesel engines in the US [24].

1.2.1 Oil Change Intervals

Harsh engine environments will eventually degrade lubricants through multiple complicated mechanisms. As discussed, these mechanisms place considerable burdens on lubricant parameters. As a result lubricants must, relatively frequently, be changed for proper engine maintenance. This creates increased hazardous waste, providing a secondary avenue for pollution from internal combustion engines. Increasing intervals is attractive, as it reduced maintenance requirements and reduces the need for hazardous waste disposal. Passenger car change intervals typically vary from 3,000 miles for short trip and severe service to over 7500 miles for less severe service, with the advent of computer based monitoring offering opportunities for better in service prediction of appropriate change intervals [25][26].

California alone is estimated to purchase 270 million gallons of new oil, and dispose of 116 million gallons of used oil each year [27]. The state estimates a doubling of oil drain interval could save $1 million per year in purchase costs alone. In situ options, using filters, have proven effective in removing particulates and maintaining basicity in some systems [28][27].

1L.2.2 Engine Durability and Failure

This work focuses on the benefits of separating the lubricating system to isolate the valve train and power cylinder lubricants. A major motivation for the consideration involved engine

durability as it relates to lubricant composition and aging. Some surveys suggest engine life may be of less concern to operators who generally experience few lubricant failures [27]. When lubrication problems do occur it tends to be due to the entrainment of combustion products and lubricant degradation. Despite this catastrophic failures can frequently result in major

maintenance delays or reduced availability and high cost. This trend has led to a general increase in oil monitoring programs. Concerns for such failures also provide barriers to extending

lubricant drain intervals, even in cases where active analysis programs are conducted.

Low viscosity -While lowering viscosity is of great benefit for fuel economy, failures may result from increased wear due to metal to metal contact. The valve train is often considered the most significant location for such contact; however the power cylinder may also suffer from

significant wear. Failures of crank assembly components during low viscosity oil tests have been reported, although with the right additive package reductions in HTHS 150 to 2.1 cP have been demonstrated [10].

Soot in valve train -Excessive soot is known to cause excessive wear and subsequent failure in valve trains. Recent studies show that soot plays a significant role in wear in low viscosity applications [10]. The author is aware of several cases. Such failure concerns are the motivation for many valve train durability tests in oil API category certification. A recent series of engine

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seizures on EPA 2010 compliant 15 liter heavy duty diesel engines, during which brass pins in roller followers were contaminated with heavy soot, showcases the ongoing nature of the

problem. In that case the oil formulator, customer, and OEM considered installation of secondary filters after finding heavy carbon deposits in the sump and in oil samples in excess of anticipated levels. In another recent case a high speed 10,000 hp marine diesel engine suffered catastrophic failure due to a failed exhaust valve stem. Of the several contributing factors were high lubricant acid levels resulting from a base number drop due to high fuel sulfur. In that case TBN fell from

9 to 2.7 in 400 hours of normal engine operation while the operator was maintaining a 1000 hour

oil change interval. High profile failures of this type are indicative of the design compromises inherent in sharing a common lubricant between the valve train and the power cylinder subsystem.

Corrosion -Valve failure may occur due to corrosion of the valve from the combustion or head side of the valve. In a dual loop lubrication system protection from valve stem corrosion may be achieved if acid levels in the lubricant are reduced. Portions of the valve head and stem will be exposed to combustion gases due to valve operation, so the engine modification will not protect against these failures. Valve train parts not exposed to the combustion chamber should receive greater protection. Deposit buildup and, or, corrosion on the valve stem or tip may develop from degraded lubricants. Wear, pitting, and scuffing are of considerable concern for cam lobe and follower wear as well. Inadequate lubrication, and debris contamination, are associated with hydraulic lift failure in those engines fitted with them.

1.3 Lubrication System Design

For the purpose of the following discussions, a brief introduction is presented. The term 'conventional system' is used to describe a lubricating system with one pump delivering lubricant to all subsystems as is typical of current engine designs for which a common system serves the valve train and crankcase subsystems. While a useful hardware configuration, it creates lubricant design tradeoffs which will be discussed in detail in this work. Recent

implementation of emissions aftertreatment systems have increased the impact of these tradeoffs

[16]. Splitting the lubrication system may decouple some of these requirements, providing

opportunities for reduced friction, emissions, and overall oil dependency. The term 'dual' or 'split' system is intended to describe an engine lubrication of atypical design for which the valve train and power cylinder subsystems have separate lubrication loops which do not interact. The

term 'split' or 'segregated' may also be used more broadly to describe any configurations for which one part of the engine's lubrication system is segregated.

1.3.1 Conventional Systems

Nearly all diesel and spark ignition engines used in automobiles today employ what is

collectively referred to in this work as a conventional lubrication system. Typically some form of

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lubricating oil at some design pressure (Figure 6). Mechanical pumps are popular due to their increased reliability over an electric pump. The flow is often controlled by a relief valve, which diverts excess flow to maintain appropriate pressures, particularly during start up when the oil temperature is low and viscosity is high.

-av -TrainTai Subbsystem

PPower Cylinder

e

Subsystem

Figure 6 Conventional (left) and dual loop (right) lubrication system configurations.

The experimental studies in this work relate to a particular small diesel engine with a crankshaft driven mechanical oil pump which draws oil from the main engine sump and delivers it to the main bearings and the valve train camshaft journal bearings and rocker arms by way of a single oil filter. The oil then lubricates other components in the engine by splashing onto them, the most significant being the pistons and rings. Oil is delivered to camshaft journals and rocker arms via pressurized oil passages. Other valve train components are lubricated by splash. A schematic of the system, modified from the Kohler shop manual, is given in Figure 7. The engine in this study was a twin cylinder with the lubrication circuit configuration.

The system described above differs from many automotive engines in that it has a belt driven camshaft. Many engines employ a chain driven camshaft, or a gear box, with lubrication draining back to the main sump.

1.3.2 Dual loop Systems

A sketch of the dual lubrication loop concept is given in Figure 6. The schematic shows a

dedicated valve train oil pump, however a multiple gear pump could also serve the same purpose and allow driving from a common shaft.

While several studies in the literature incorporate standalone cylinder heads and motoring tear down tests, there are few studies or proposals using split systems either as test stands or as practical systems in the automotive and off road industry [29].

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ON 4 F~ter Camshaft Pressure OB Pump Crankshaft

Figure 7 Lubricant system for 3 cylinder Kohler KDW 1003 (turbocharger removed) [30][31]. A similar configuration, the KDW 702, a twin cylinder naturally aspirated diesel engine, was used in this study. The marine industry operates large cross head two stroke engines. These engines have separate lubrication for the cylinder and crankcase, leading to different oil formulations as a result of the significant difference in Sulfur content in the crank case as opposed to the head [32][33]. In the case of cross head engines, the separation is possible due to the presence of a sealing gland [29].

A recent study by Nattrass used a split system to study friction effects on a spark ignition engine

[34]. Schilling describes a special Renault single cylinder test engine with separate loops, but noted it was not yet used for oil testing [35]. Mufti et al recently used an independent valve train lubricant system on a single cylinder SI engine and instrumented the camshaft pulley with torque sensors in a manner similar to the current study for the purpose of observing valve train friction

[36]. Similar works are not available on diesel or spark ignition engines based on searches of SAE and STLE literature.

Some patents were issued over the last 30 years regarding the development of split systems; however they are not incorporated in practice in any automotive fleet. There are some patents related to separate valve train lubrication systems, the most significant were granted to Southwest Research Institute [37] and Nissan [38].

Additional patents exist for lubricant formulations specifically tailored to an engine with separate lubricating subsystems. Lubrizol patented separate oils for the valve train and power cylinder systems on a spark ignition engine explicitly [39]. They also patented a low sulfur consumable lubricant for the power cylinder subsystem of CI or SI engines [40].

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In this study a dual loop system is developed using a small 16 hp diesel engine, the Kohler KDW

702. The oil drains from the head are plugged and a separate oil pump and sump are used for the

cylinder head. The valve train subsystem consists of the camshaft, rocker arms, valve stems, and components in the cylinder head. The rest of the engine, including the piston and rings, and crankshaft, encompass the power cylinder subsystem. The system is described in greater detail in Chapter 2.

1.4 Lubricant Formulation

Typical automotive lubricants consist of base oil and an additive package. The base oil accounts for roughly 75-90% of an engine oil formulation by mass. Viscosity modifiers, if used, account for approximately 10% of the total formulation. The rest of the additive formulation consists of the detergent inhibitor (DI) package and other additives, with dispersants accounting for 6%, antiwear 1.5%, and detergents 3.5% [41]. In general additive impacts, other than the viscosity changes from viscosity modifiers, have a limited effect on overall fuel economy as compared to viscosity changes [42]. This is likely due to the significantly greater portion of friction that is attributed to hydrodynamic losses in the power cylinder system as will be discussed. Additive optimization is expected to have greater impact on valve train friction [411.

1.4.1 Base Stock

Base stocks are categorized by their American Petroleum Institute (API) groups. They are defined by API 1509, Appendix E [43]. A summary is provided in the following table.

Table 1 API Base Stock Group Definitions

API Group Saturates (%) Sulfur (%) Viscosity Index Notes

1 <90 >0.03 80-120 Mineral base

I >90 <0.03 80-120 Mineral base

III >90 <0.03 >120 Mineral base

IV N/A N/A N/A All PAO (full synthetic,

polyalphaolefins)

V N/A N/A N/A Any base stock not

included in Group l-IV

The first four groups are currently used to varying degrees in automotive applications. Group I and II are the most common base stocks, and are the least expensive. Group III base stocks are becoming more widely used in industry. Group II and III base oil demand is anticipated to grow

by 4 and 10 percent annually, respectively, through 2017 [8]. The majority of Group III oil

demand is in the engine oil industry. Group II lubricants are easier to process and have higher

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saturates and solvency, making them useful for soot dispersancy. They are key components in 15W40 lubricants, which still comprise 80 percent of the heavy duty diesel market [8]. The primary advantage of Group III base stock is the higher viscosity index (VI), making it useful for formulations such as 5W30 [8]. In addition, synthetic lubricants have better thermal and

oxidative stability. Group IV PAO's are pure synthetics; however their high cost makes them less prevalent than the lower groups. Estimates place the cost of PAO base stocks at 2.5-3 times those of Group I and II [44].

Group V base oils consist of all other stock not included in the other categories. They include polyalkylene glycol (PAG) and esters. Group V base stocks, while exhibiting better temperature

viscosity characteristics, may not be suitable for current engine applications due to chemical reactions with materials. For instance, while providing better VI, the impact on seals from PAG oils are not well understood. Recent developments with ionic liquid additives indicates 2% improvements in fuel economy may be realized [45].

1.4.2 Additives

An exhaustive description of additive technology is beyond the scope of this work. Only a brief description will be provided here. Several useful primers are available which may be of interest to readers [32][25][33]. Additive technology continues to advance. The majority of additive types were identified in the 1930's and 40's [46]. Despite this many additive interactions are yet to be understood [47].

Viscosity Modifiers - Viscosity modifiers, or Viscosity Index (VI) Improvers, are often required to reduce viscosity at low temperature regimes in Group I, II, and III base oils. They allow lower viscosity base oils with better temperature viscosity characteristics to be used and were first introduced to reduce low temperature viscosities for better cold starting [48]. Dilute solutions of the polymers increase the viscosity of the base oil, raising it at all temperatures without

significantly changing the viscosity temperature slope [49]. Dilute polymer solution viscosities may be estimated using the Higgins Equation with viscosity increasing as molecular weight increases [50]. While most base oils exhibit molecular weights up to 400, viscosity modifiers are much greater, ranging from 10,000 to 20,000 [51]. The most common types are

polyisobutylenes, olefin copolymers (OCP), and polymethacrylates (PMA) [52], although several other types are also available [32]. OCP are generally preferred due to low cost relative to their performance. Both OCP and PMA have an added advantage in that they may be formulated with dispersant characteristics [32].

The degree to which viscosity is increased at different temperatures varies depending on the polymer type, with most, including OCP, resulting in relatively greater increases at lower temperatures than at higher temperatures. For the base oil with 10% OCP shown in Figure 8 the viscosity was increased by 74% at 20'C but only 61% at 200'C. Polyalkyl methacrylates (PMA) are an exception, with a relative viscosity increase as temperature increases [48]. Ver Strate shows a gain of approximately 15% at 20'C and 140% at 160'C using PMA to create a 5W20