AN EXPERIMENTAL STUDY OF POWER LOSSES OF AN AUTOMOTIVE MANUAL TRANSMISSION. A Thesis. The Degree of Master of Science in the

Transcription

1 AN EXPERIMENTAL STUDY OF POWER LOSSES OF AN AUTOMOTIVE MANUAL TRANSMISSION A Thesis Presented Partial Fulfillment of the Requirements for The Degree of Master of Science the Graduate School of the Ohio State University By Timothy A. Szweda, B.S. * * * * * The Ohio State University 2008 Masters Examation Committee: Dr. Ahmet Kahraman Approved by Dr. James Schmiedeler Advisor Department of Mechanical Engeerg

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3 ABSTRACT In this study, the fluence of a variety of operatg conditions on the power losses and efficiency of an automotive manual transmission was vestigated experimentally. An experimental methodology was developed to measure power losses of a manual transmission under both loaded and unloaded conditions while all operation parameters were controlled tightly. A set of fixtures and strumentation were designed and implemented to apply the experimental methodology to a six-speed, manual transmission from a front-wheel-drive passenger vehicle. Experimental parametric studies were performed to quantify the fluence of operatg conditions cludg speed, nomal torque transmitted, oil temperature and oil volume on load-dependent (mechanical) and load-dependent (sp) power losses of the transmission. Analysis of the power loss data revealed that all four of these parameters fluenced the components of the transmission power loss significantly, and specific conclusions were drawn order to aid attempts to crease overall transmission efficiency. The experimental database formed as a result of this study is extensive so as to allow a complete validation of transmission power loss models. ii

4 Dedicated to my parents. iii

5 ACKNOWLEDGMENTS I would like to thank Dr. Ahmet Kahraman for all of his leadership and guidance; from encouragg me to stay for my master s degree and vitg me to work on this project, to his thorough review of my thesis. I would also like to thank all of the GearLab students and staff for their support, especially Sam Shon for his valuable assistance my experimental work and Satya Seetharaman for his contributions to all aspects of this project. Tremendous thanks to Gary Gardener, who displayed credible patience with me addition to his expertise machg. Thanks also go to Dr. James Schmiedeler for his willgness to serve on my master s examation committee, and to General Motors Powertra Europe for their sponsorship of this project. I would also like to thank my family for their support and encouragement throughout my life. I would not be where I am today without the contual guidance of my parents, and the focus on learng that they stilled me from the very begng. Thanks also go to all of my friends, especially my roommates, who have provided me not only with support, but with the stress relief and fun that I needed to survive the past five years. Fally, I would like to thank Lara Wygle for her encouragement, support, and love. iv

6 VITA June 7, 1985 Born Akron, Ohio June 2004 Sept Mechanical Engeerg Intern, General Electric Healthcare Aurora, Ohio June 2005 Sept Mechanical Engeerg Intern, General Electric Healthcare Milwaukee, Wiscons March 2006 June 2006 Undergraduate Student Assistant The Ohio State University Columbus, Ohio June 2006 Sept Mechanical Engeerg Intern, General Electric Aviation Ccnati, Ohio Sept June Undergraduate Research Assistant The Ohio State University Columbus, Ohio June 2007 B.S Mechanical Engeerg The Ohio State University Columbus, Ohio June 2007 August Graduate Research Associate The Ohio State University Columbus, Ohio Major Field: Mechanical Engeerg FIELDS OF STUDY v

7 TABLE OF CONTENTS Page Abstract....ii Dedication...iii Acknowledgments..iv Vita...v Table of Contents vi List of Figures...viii List of Tables...xiii Nomenclature...xiv Chapters: 1 Introduction Background and Motivation Literature Survey Thesis Objectives Thesis Outle..7 2 Experimental Test Methodology Description of Transmission Efficiency Test Set-up Test Procedure...19 vi

8 2.3 Test Transmission and Test Matrix Accuracy and Repeatability Transmission Efficiency Test Results Calculation of Transmission Efficiency Transmission Efficiency Test Results Influence of Input Speed Ω Influence of Oil Temperature θ Influence of Nomal Input Torque T Influence of Oil Volume V oil Summary and Conclusions Summary Conclusions Recommendations for Future Work..78 Bibliography 80 Appendix A Design Specifications of Major Mache Elements...82 vii

9 LIST OF FIGURES Figure Page Figure 2.1. A view of the transmission dynamometer set-up used this study with the safety guards on the output side removed Figure 2.2. A schematic of the lay-out showg the components of the set-up Figure 2.3. Close-up view of (a) the put side and (b) the output side torquemeters Figure 2.4. Optical tachometer used to measure rotational speed of the put shaft Figure 2.5. Type K thermocouple serted to transmission dra plug to measure transmission oil temperature Figure 2.6. A screen capture of DASYLab user terface Figure 2.7. Cutaway view of the 6-speed manual transmission used this study [1] Figure 2.8. Power flow paths at each gear stage of the example manual transmission [1] Figure 2.9. A schematic side view of the transmission showg the static oil levels at various oil volumes Figure Blended thermal images of the transmission at oil temperatures of (a) 30, (b) 50, and (c) 80 C...26 Figure A typical variation of the measured stantaneous put speed Ω with time at a set put speed of 2000 rpm Figure Typical variation of oil temperature θ with time Figure Typical variation of measured (a) T and (b) out T with time at a set value of 50 Nm durg operation at 6 th gear stage viii

10 Figure Repeatability of the test set-up and the measurement system through a comparison of the results of two tests performed under the same conditions of T = 50 Nm, Ω = 4000 rpm, V oil = 3.1 liters, and θ=80 C...32 Figure 3.1. A typical measured variation of the total power loss P T and transmission oil temperature θ a function of time. T = 50 Nm, Ω = 2000 rpm and V = 2.0 liters oil Figure 3.2. Measured (a) P T, (b) P s and (c) P m at T = 50 Nm, θ= 80 C and V = 3.1 liters oil Figure 3.2. Contued. Measured (a) P T, (b) P s and (c) P m at T = 50 Nm, θ=80 C and V oil = 3.1 liters Figure 3.3. Comparison of measured variation of P T, P s and P m with Ω at (a) 2 nd, (b) 3 rd, (c) 4 th, (d) 5 th and (e) 6 th gear stages. T = 50 Nm, θ=80 C and V oil = 3.1 liters Figure 3.3. Contued. Comparison of measured variation of P T, P s and P m with Ω at (a) 2 nd, (b) 3 rd, (c) 4 th, (d) 5 th and (e) 6 th gear stages. T = 50 Nm, θ= 80 C and V = 3.1 liters oil Figure 3.3. Contued. Comparison of measured variation of P T, P s and P m with Ω at (a) 2 nd, (b) 3 rd, (c) 4 th, (d) 5 th and (e) 6 th gear stages. T = 50 Nm, θ= 80 C and V = 3.1 liters oil Figure 3.4. Measured (a) η T and (b) η m at T = 50 Nm, θ= 80 C and V oil = 3.1 liters...44 Figure 3.5. Measured fluence of θ on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 2, 000 rpm and V oil = 3.1 liters ix

11 Figure 3.5. Contued. Measured fluence of θ on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 2, 000 rpm and V = 3.1 liters Figure 3.6. Measured fluence of θ on (a) η T and (b) η m at T = 50 Nm, Ω = 2,000 rpm and V = 3.1 liters oil Figure 3.7. Measured fluence of θ on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 4, 000 rpm and V oil = 3.1 liters Figure 3.7. Contued. Measured fluence of θ on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 4, 000 rpm and V = 3.1 liters Figure 3.8. Measured fluence of θ on (a) η T and (b) η m at T = 50 Nm, Ω = 4,000 rpm and V = 3.1 liters oil Figure 3.9. Measured fluence of T on (a) P T and (b) P m at θ= 80 C, Ω = 2,000 rpm and V = 3.1 liters oil Figure Measured fluence of T on (a) η T and (b) η m at θ= 80 C, Ω = 2,000 rpm and V oil = 3.1 liters Figure Measured fluence of T on (a) P T and (b) P m at θ= 80 C, Ω = 4,000 rpm and V = 3.1 liters oil Figure Measured fluence of T on (a) η T and (b) η m at θ= 80 C, Ω = 4,000 rpm and V = 3.1 liters oil Figure Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 2, 000 rpm and θ= 30 C...60 oil oil x

12 Figure Contued. Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 2, 000 rpm and θ= 30 C...61 Figure Measured fluence of V oil on (a) η T and (b) η m at T = 50 Nm, Ω = 2,000 rpm and θ= 30 C...63 Figure Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 2, 000 rpm and θ= 80 C...64 Figure Contued. Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 2, 000 rpm and θ= 80 C...65 Figure Measured fluence of V oil on (a) η T and (b) η m at T = 50 Nm, Ω = 2,000 rpm and θ= 80 C...66 Figure Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 4, 000 rpm and θ= 30 C...68 Figure Contued. Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 4, 000 rpm and θ= 30 C...69 Figure Measured fluence of V oil on (a) η T and (b) η m at T = 50 Nm, Ω = 4,000 rpm and θ= 30 C...70 Figure Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 4, 000 rpm and θ= 80 C...71 Figure Contued. Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 4, 000 rpm and θ= 80 C xi

13 Figure Measured fluence of V oil on (a) η T and (b) η m at T = 50 Nm, Ω = 4,000 rpm and θ= 80 C...73 Figure Combed fluence of V oil, Ω and θ on (a) P T and (b) η T at 6 th gear range and T = 50 Nm Figure A.1. Input and output torque-meter specifications (source: Figure A.2. Input and output torque-meter dimensions (source: Figure A.3. Flexible couplg specifications (source: Figure A.4. Flexible couplg dimensions (source: Figure A.5. Ma put side support plate Figure A.6. Ma output side support plate Figure A.7. Transmission mountg plate Figure A.8. Input side torque-meter lower adapter plate Figure A.9. Input side torque-meter upper adapter plate Figure A.10. Transmission output shaft adapter Figure A.11. Eddy current dynamometer support Figure A.12. Speed creaser support Figure A.13. Output side safety cover xii

14 LIST OF TABLES Table Page 2.1. The test matrix used this study. 24 xiii

15 NOMENCLATURE Symbol P T T V Γ η θ Ω ω Description Power [kw] Torque [Nm] Nomal torque [Nm] Oil volume [liters] Gear ratio Efficiency Oil temperature [ C] Rotational speed [rpm] Rotational speed [rad/s] Subscripts fd i m oil out s sgi T Fal drive Gear stage Input Mechanical power loss Manual transmission fluid Output Sp power loss Speed gear pair Total power loss xiv

16 CHAPTER 1 INTRODUCTION 1.1 Background and Motivation The loomg energy crisis and environmental concerns regards to global warmg and air quality have recently placed great emphasis on fuel economy performance and gas and particulate emissions of passenger and commercial vehicles. Both emissions and fuel consumption of a vehicle are fluenced largely by the efficiency of the power tra of the vehicle. Power losses of a power tra can be traced back to the herent losses of the enge generatg the power and the losses that occur durg transmission of power to the wheels through the drive tra. The transmission can be identified as the major component of the drive tra not only terms of its contributions to the power losses, but also the potential it presents for improvg overall power tra efficiency. This renewed terest more efficient transmissions promoted the need for a fundamental understandg of the factors causg power losses before any logical and cost-effective steps could be taken to mimize them. It is safe to categorize the power losses of a transmission to two groups. The first group cludes all losses associated 1

17 with the transmission of torque. These load-dependent (or mechanical) losses are all duced by friction at contactg terfaces of the transmission. Gear meshes and the rollg element beargs provide multiple contacts, contributg significantly to the mechanical losses of the transmission [1]. The other group of power losses is associated mostly with the teractions between the surroundg medium (oil, air or a mixture of the two) with the rotatg components such as gears and beargs [2, 3]. These sp power losses are dependent of load transmitted and are dictated mostly by factors such as effective oil levels, rotational speeds, transmission temperature and the geometry of the transmission housg. This study aims at quantifyg experimentally (i) the dividual contributions of sp and mechanical losses and (ii) the fluence of several operatg parameters on the total power losses and overall efficiency of a manual transmission. 1.2 Literature Survey A significant number of published studies are available on the power losses and efficiency of dividual gear pairs and gear tras, as has been reviewed papers by Mart [4], Yada [5], Li and Seireg [6], Seetharaman and Kahraman [2], and Seetharaman, et al. [3]. However, literature on the efficiency and power losses of manual transmissions is rather sparse. Hegartner and Mba [7] proposed a mathematical model that predicted the power losses of helical gear transmissions and provided experiments for a validation of the model. They divided power losses to speed- and load-dependent losses. The speed-dependent losses were further divided to wdage losses, churng losses, bearg churng losses and seal losses. Wdage losses were suggested to be 2

18 caused by the fe mist of oil droplets that is flung off of the rotatg gears to cause frictional resistance and turbulence with the gearbox. They stated that the oil churng losses are caused by the entrapment of the lubricant with the gear mesh. The loaddependent losses were further divided to slidg friction losses, rollg friction losses, and bearg losses. Slidg friction losses were determed by the stantaneous slidg velocity and the friction force, which is dependant on the normal tooth load and the stantaneous coefficient of friction. Rollg friction losses were determed by the stantaneous rollg velocity and the stantaneous lubricant film thickness. A set of experiments usg an example set-up was performed at various load levels at constant speed. They concluded that creasg the load at a constant speed creased the slidg losses and slightly decreased the rollg friction losses, while the wdage losses remaed constant. Changenet, et al. [8] provided a theoretical and experimental comparison of power losses a six-speed manual gearbox. In this study, power losses were aga separated to no-load and load-dependent components. Load-dependent losses considered cluded friction at matg teeth and rollg element beargs. No-load losses considered cluded oil churng and oil shearg journal beargs and synchronizer cones, while wdage losses were not cluded. Load and temperature effects were vestigated to conclude that the transmission efficiency is higher at elevated temperatures and higher loads. Handschuh and Kilma [9] performed an experimental and analytical comparison of the efficiency of high-speed helical gear tras used a tilt-rotor aircraft. In the 3

19 experimental portion of the study, two identical gearboxes were arranged a closed-loop power circulation arrangement, and the load and speed were considered as the ma variables. They concluded based on the experimental and analytical results that higher gear rotational speeds creased the power losses of each gear tra significantly. Conversely, they stated that the torque transmitted by the gear tra has a relatively significant effect under most conditions, especially when it is compared to the fluence of the rotational speed. In this set-up, air wdage power loss was stated to be comparable to gear meshg losses at high load and high speed conditions. Meanwhile, the wdage power loss was reported to domate the overall losses and at light loads and higher speeds. Dongen [10] performed a study usg both manual and automatic transmissions to quantify the fluence of speed and load on transmission efficiency. Loaded and unloaded tests were performed at various load and speed levels while the temperature of the transmission was kept constant. He concluded that efficiency creases with load, and that the maximum efficiency that could be reached by a manual transmission was about 96%. He also identified oil churng, bearg friction and load dependent gear mesh losses as the ma contributors to manual transmission power losses. Barzaghi, et al. [11] performed a test which two manual transmissions were used the same test rig, one as a speed reducer and one as an overdrive. An average value of efficiency was estimated based on the torque measurements. They reported load as the most important parameter fluencg efficiency, with lower efficiency levels at reduced load conditions. They also concluded that lubricants with anti-friction additives result greater efficiency, especially at higher loads. It was also noted their study that 4

20 the actual oil temperature with a transmission operatg under conditions representg those a vehicle can crease to levels as high as 110 C. In another manual transmission efficiency study, Greenbaum, et al. [12] vestigated the fluence of the clation angle of the transmission and the transmission oil temperature. The transmissions that were tested varied widely efficiency from %. It was found that creasg the clation angle decreased efficiency, with a difference of 5 degrees of clation (which is a typical variation, based on how the same transmission can be mounted different vehicle models) causg as much as 1% difference transmission efficiency. They also observed that the efficiency goes up significantly with creased oil temperature. For stance, they measured a 6% crease efficiency of the transmission when the oil temperature is creased from 50 to 75 C. They also concluded that torque rather than speed is the predomant factor affectg transmission efficiency gear stages other than direct drive. Several other studies focused on corporatg efficiency requirements early stages of design by implementg design changes or methods durg the conception of the transmission. In one such study, Kluger, et al. [13] focused on drastic design changes to a transmission to achieve efficiency improvements under actual duty cycle requirements of passenger vehicles. They suggested several specific methods of improvg manual transmission efficiency based on these duty cycle requirements, cludg adoptg a lube pump and jet lubrication method to reduce oil churng losses and addg a synchronizer on the put shaft to disengage the layshaft when 4 th gear (the direct drive gear) to further reduce wdage losses. They also proposed a dual layshaft transmission which 5 th gear would get its own layshaft, which would reduce 5

21 wdage from other gears and shafts that would be rotatg when not use. Another study of the same kd [14] measured the average efficiency of a five-speed manual transmission to be 96.2%. Several specific methods to improve the efficiency of this transmission were suggested. One proposal was to crease the layshaft speed, which they claimed would help to reduce torque-dependant losses at the expense of the less important speed losses. Others were to reduce the castg imperfections on the side of the transmission case to reduce wdage losses and to add an oil-sheddg dry film lubricant to ner case surface. It was also proposed to change from steel to ceramic ball beargs to reduce parasitic losses. It was found that corporatg all of these improvements would result a modest 0.5% crease transmission efficiency. 1.3 Thesis Objectives The existg body of research on manual transmission efficiency fails to provide a consistent picture of the transmission efficiency as many conclusions from different studies conflict with each other. None of these studies is extensive enough to provide a comprehensive database that cludes all key operatg parameters. Neither the accuracy and sensitivity of the fixturg and measurement systems nor the repeatability of the measurements were reported these studies. More importantly, these studies fail to report the critical details of the test conditions as well as parameters of the test transmissions, makg them unsuitable for quantitative model validation efforts. Accordgly, this study aims at conductg manual transmission experiments an 6

22 attempt to void some of these shortcomgs of the previous studies. Specific objectives of this study are as follows: Develop an experimental methodology to measure power losses of a manual transmission under tightly controlled loaded and unloaded operatg conditions. Design and implement a set of fixtures and strumentation to apply the experimental methodology to a specific six-speed manual transmission. Demonstrate the accuracy and the repeatability of the measurements. Perform experimental parametric studies to quantify the fluence of operatg conditions, cludg speed, load, oil temperature and oil volume on mechanical (load dependent) and sp (load dependent) power losses. Identify the key parameters fluencg the components of the transmission power loss order to arrive at guideles on how to crease overall transmission efficiency. 1.4 Thesis Outle Chapter 2 provides a description of the experimental test methodology that was developed this study. It begs with a description of the test mache, which covers the mache components and layout as well as the software and data gatherg. The test procedure is explaed, and the test transmission and test matrix are described. Fally, the accuracy and repeatability of the test set-up are vestigated. Chapter 3 details the results of the experimental study. The method that is used to calculate transmission efficiency from the raw data is explaed, and detailed results are 7

23 presented. The data are presented various formats order to demonstrate the effect of oil volume, oil temperature, speed and torque. The fluences of these parameters are presented terms of the separate components (sp and mechanical) of power loss, as well as overall power loss and efficiency. Chapter 4 provides a summary of the results and detailed conclusions that can be drawn from them. The key parameters fluencg the components of power loss are identified, and potential ways for creasg the overall transmission efficiency are suggested. Fally, recommendations for future work are listed. 8

24 CHAPTER 2 EXPERIMENTAL TEST METHODOLOGY 2.1 Description of Transmission Efficiency Test Set-up A special-purpose test set-up was developed this study to allow a direct measurement of the power loss of an example automotive manual transmission under varyg torque, speed, oil temperature and oil volume conditions at each gear range. An image of the test set-up is shown Figure 2.1, and a schematic of the layout with each component labeled is provided Figure 2.2. In this arrangement, the manual transmission was held by a massive bracket that provides an terface that is identical to that of the tended ternal combustion enge a vehicle. A set of bolts were used to hold the transmission rigidly its place on the holdg bracket such a way that all the axes of the transmission are horizontal and the tilt angle about the put shaft axis is the same as that of the actual application for this transmission. At the put side, a 110 kw DC motor was used to drive the transmission at any desired speed-torque conditions with the limits of the drive unit. A slender shaft was connected to the drive motor via a rigid couplg and to the put torque-meter via a 6-9

25 10 Figure 2.1. A view of the transmission dynamometer set-up used this study with the safety guards on the output side removed. 10

26 Lebow 1703 Input T/M 100 Nm DC Drive Motor 110 kw Sgle Flex, 6-bolt couplgs Lebow 1706 Output T/M 1500 Nm Speed Increaser 6-speed Manual 76:47 Ratio Transmission Sgle Flex, 6-bolt couplgs Absorber 400 kw 11 Figure 2.2. A schematic of the lay-out showg the components of the set-up. 11

27 bolt, sgle-disc, flexible couplg (Lovejoy SU 90-6). The put torque-meter (Lebow 1703) shown Figure 2.3(a) had a maximum allowable speed of 13,500 rpm and a maximum torque level of 100 Nm with an accuracy of 0.25% of this maximum torque value. This corresponded to accuracy up to 0.25 Nm that was deemed sufficient for this application. This transducer was used to measure put torque values T applied to the transmission by the DC drive. The torque-meter was mounted on a pedestal plate between two flexible couplgs so that it could be removed for matenance purposes by simply disengagg the couplgs. One of the flanges of the couplg between the put torque-meter and the transmission put shaft was spled to terface with the put shaft as it is vehicle operation. The flexible couplg flanges that were connected to the torque-meter were both keyed and shrk-fit onto the torque-meter shaft to ensure no slip. On the output side of the transmission, another precision torque-meter was connected to the output shaft at the opposite side of the put shaft by a specially modified sgle-disc, 6-bolt, flexible couplg (Lovejoy SU 158-6), which connects to the output torque-meter. This torque-meter (Lebow 1706), shown Figure 2.3(b), has a maximum allowable speed of 7,900 rpm and a maximum torque range of 1,500 Nm with an accuracy of 0.25% of the maximum torque value. This transducer measures output torque values ( T out ) experienced by the transmission under loaded operatg conditions. On the other side of the output torque-meter, another flexible couplg and a long shaft supported by a pair of rollg-element beargs provide the connection to a speedcreaser (76:47 ratio) gearbox, that is connected to an eddy-current load brake dynamometer (400 kw), both of which are water cooled. Sce the speed reduction of 12

28 (a) (b) Figure 2.3. Close-up view of (a) the put side and (b) the output side torque-meters. 13

29 the example transmission first gear was excess of 13:1, this speed creaser was required to boost the rotational speed of the brake dynamometer so that the required output torque could be generated. All of the components of the output-side drive tra between the transmission and the speed creaser gearbox were supported by a sgle pedestal plate. In this arrangement, capability to disengage the couplg between the transmission and the output torque-meter was provided so that no-load (sp) tests could be performed with the output side disconnected. The components of the put side drive tra and the put shaft axis of the transmission were aligned to the axis of the DC motor through a rotational laser alignment system. The output drive tra was also aligned the same way. This ensured that both the horizontal and vertical misalignments were mimal. Further, the flexible couplgs on both sides of each torque-meter provided addition means to prevent radial loadg of the torque-meters. A measurement system that tightly controls the put speed Ω ( Ω = 60 ω 2π rpm) and the put torque T was corporated to the experimental set-up. The put speed is measured usg an optical tachometer, which displays the current speed to the test operator and at the same time provides feedback to allow a manual adjustment by turng a potentiometer. In this way, the put speed was constantly monitored and mataed with ± 1% of the desired set value. Figure 2.4 shows the optical tachometer directed toward a taped section of the put shaft. 14

30 Figure 2.4. Optical tachometer used to measure rotational speed of the put shaft. 15

31 The DC drive motor was capable of reachg put speeds ( Ω ) up to 4500 rpm all gear ratios. The load brake dynamometer was connected to a variable power supply that was capable of supplyg 60 volts and 4 amps to the dynamometer. This dynamometer set-up was able to provide put torque values up to 100 Nm, except the first gear range where it was limited to 30 Nm. Usg this system, the put torque was manually adjusted and mataed with ± 2 Nm of the desired set value durg each test. The temperature of the transmission was measured contuously by usg a type- K thermocouple with a lear range of 0 to 100 C. The thermocouple was serted to the bottom dra hole of the transmission, as shown Figure 2.5. At this lowest location of the transmission, likelihood of havg direct oil contact to the thermocouple is the highest, which creases the chance of the measured temperature value representg the true oil temperature with the transmission. Durg these tests, no attempt was made to control the temperature of the transmission. Instead, the temperature was allowed to vary as the transmission operates, and a time history of the temperature data was obtaed. This allowed extraction of power losses at certa desired temperatures through processg of the data. The outputs of the speed tachometer, thermocouple and torque-meters were all sent to a PC runng DASYLab data acquisition software. A 20-pot runng average was used to average the put signals, and the data was converted from the put voltage to the proper engeerg units for each sensor. The data was saved to a text file, which 16

32 Figure 2.5. Type K thermocouple serted to transmission dra plug to measure transmission oil temperature. 17

33 Figure 2.6. A screen capture of DASYLab user terface. 18

34 was then imported to a spreadsheet for further processg. A screen capture showg the DASYLab user terface can be seen Figure Test Procedure Before any particular test was itiated, a certa volume of oil was measured and poured to the transmission, which was emptied completely after the previous test. When a test was performed to measure transmission power losses under loaded conditions, the entire set-up shown Figures 2.1 and 2.2 was used with both put and output torque-meters active and connected to the transmission. In case of unloaded tests to measure sp power loss, the couplg to the left of the output torque-meter, shown Figure 2.2, was disengaged so that T out = 0 and hence T = T loss. In each case, the dynamometer was then activated and brought up to the desired speed. For loaded tests, an operator-defed nomal put torque T was applied by adjustg the voltage sent to the eddy current brake dynamometer. When each of these preset parameters was achieved, data collection DASYLab was itiated. The test was run for a certa period of time to allow the temperature of the transmission to climb to its steady state, on its way passg through the discrete temperature values of terest with this range. For temperature values exceedg this steady state level, the test was started with heated oil and allowed to cool to the same steady-state temperature level. This simple procedure allowed measurements with the temperature range of 20 to 90 C. 19

35 2.3 Test Transmission and Test Matrix The example test transmission considered this study was a six-speed manual transmission used front-wheel-drive passenger car applications. It was provided by the sponsor of this project assembled form with a locked differential. The locked differential prevents any relative motion between the two output shafts (deactivates the differential completely) so that the output torque can be measured by usg only one of the output shafts. Figure 2.7 shows a cutaway illustration of this transmission that consists of four shafts (an put shaft, two transfer shafts and a differential shaft) that are supported by pairs of tapered roller beargs to hold ne pairs of gears to achieve six distct forward speed ratios (gear stages) as well a reverse gear ratio. Figure 2.8 shows the power flow paths associated with each gear stage. The example transmission used this study has the overall speed ratios Γ i =Ω Ω out of First gear: Γ 1 = Second gear: Γ 2 = 7.23 Third gear: Γ 3 = 4.68 Fourth gear: Γ 4 = 3.38 Fifth gear: Γ 5 = 2.68 Sixth gear: Γ 6 = 2.21 It is noted from the power flow paths shown Figure 2.8 for the forward gear stages that there are only two loaded gear pairs for each stage. Hence, the overall ratio at a given gear stage i is the product of the gear ratio Γ sgi of the loaded speed gear pair and the fal drive ratio Γ fd, i.e. Γ i =Γ sgi Γ fd. The fal drive ratio of this transmission is Γ = fd 20

36 Figure 2.7. Cutaway view of the 6-speed manual transmission used this study [1]. 21

37 Figure 2.8. Power flow paths at each gear stage of the example manual transmission [1]. 22

38 Table 2.1 shows the experimental test matrix that was considered this study to vestigate the fluence of the various operatg conditions on the transmission power loss. The test matrix cluded (i) three oil volume V oil values, (ii) three temperature θ levels, (iii) three nomal put torque T values, (iv) four put speed Ω values, and (v) all six forward gear stages. It covered various combations of these five variables as well to vestigate their combed fluences. Oil volume values of V oil = 1.2, 2.0 and 3.1 liters were chosen to represent low, medium and high static oil levels of the transmission. A schematic of the side view of the transmission is given Figure 2.9 with the shaft axes and the static oil levels marked for these V oil values. Similarly, the discrete oil temperatures of θ = 30, 50 and 80 C represent typical values with the operatg temperature range for the transmission. The manual transmission fluid used as oil these tests had kematic viscosity values of 42, 20, and 10 cst at 30, 50 and 80 C, respectively. While power losses at much colder temperatures are also of terest, such cold-temperature tests were not cluded the test matrix sce the test facility could not achieve such conditions. Figure 2.10 shows thermal images of the exterior of the transmission at these three oil temperature levels of terest. These thermal images dicate that the thermocouple location shown Figure 2.7 represents the highest oil temperatures observed on the outside surface of the transmission housg. Tests under loaded conditions were performed at T = 50 and 100 Nm. In addition, for each of these tests a correspondg test at the same speed, temperature and 23

39 Test V oil Number [liters] 1 θ [ C] T [Nm] Ω Gear Range [rpm] , , , , , , , , , , , , , Table 2.1. The test matrix used this study 24

40 upper transfer shaft put shaft 3.1 liter 2.0 liter 1.2 liter lower transfer shaft output shaft Figure 2.9. A schematic side view of the transmission showg the static oil levels at various oil volumes. 25

41 (a) (b) (c) Figure Blended thermal images of the transmission at oil temperatures of (a) 30, (b) 50, and (c) 80 C 26

42 oil level conditions was also performed with T = 0 so that the difference between the loaded and the unloaded tests could be used for determg the load-dependent mechanical losses. The unloaded tests refer to the case where the transmission is disconnected from the output side completely, such that T = 0 and the measured T represents the torque loss. These tests were carried out at put speeds rangg from Ω = 1000 to 4000 rpm for the available gear stages. Loaded tests were not performed the first gear stage because the test set up was not able to apply sufficient put torque this case. With all of these parameter values, the test matrix shown Table 2.1 consisted of a total of 124 tests grouped to 15 sets of experiments, each set representg a different temperature, oil level and put torque condition. The tests with different oil levels were performed with at least two different temperature levels to provide a direct comparison of the trends at different oil temperature conditions. 2.4 Accuracy and Repeatability In this section, typical results for measured and controlled test parameters will be illustrated for a trial test (at T = 50 Nm, Ω = 2000 rpm, V = 1.2 liters and θ=80 C) to demonstrate the fidelity of the measured quantities. Figure 2.11 shows the variation of the measured put speed with around 1% of the desired value of 2000 rpm. Figure 2.12 shows the variation of 27 Ω with time. The maximum variation Ω is oil

43 Ω [rpm] Time [m] Figure A typical variation of the measured stantaneous put speed at a set put speed of 2000 rpm. Ω with time 28

44 θ [ C] Time [m] Figure Typical variation of oil temperature θ with time. 29

45 the measured oil temperature θ over this same time period. The temperature creases a nonlear fashion along a smooth curve, startg from the itial temperature value of around 35 C and reachg just over 80 C after 35 mutes of operation of the transmission at these conditions. Figures 2.13(a) and (b), respectively, show the variation of the measured T and T out values over this time period. The two sharp creases at 8and 19 mutes are due to manual creases the applied torque by the operator to compensate for the change the temperature values. Such adjustments were no longer needed once the transmission temperature reached its steady state. In order to ensure that no bias or drift was experienced over the course of the test program, frequent repeatability tests were performed. Results of such a repeatability test are shown Figure 2.14 at T = 50 Nm, Ω = 4, 000 rpm, V oil = 3.1 liters and θ=80 C at gear stages two through six. The compared quantity is the measured power loss P T whose calculation will be described later Chapter 3. The greatest discrepancy between two test pots is less than 0.12 kw, with the average difference between tests beg less than 2%. This shows that a reasonable level of repeatability was achieved even though these two tests were performed several weeks apart. 30

46 60 ( a) 50 T [Nm] ( b) T out [Nm] Time [m] Figure Typical variation of measured (a) T and (b) T out with time at a set value of 50 Nm durg operation at 6 th gear stage. 31

47 5 4 Test 1 Test 2 T P [kw] Gear Stage Figure Repeatability of the test set-up and the measurement system through a comparison of the results of two tests performed under the same conditions of T = 50 Nm, Ω = 4000 rpm, V = 3.1 liters, and θ= 80 C. oil 32

48 CHAPTER 3 TRANSMISSION EFFICIENCY TEST RESULTS 3.1 Calculation of Transmission Efficiency level T When the torque measurements are taken under a user-defed nomal put load, the two torque transducers measure the stantaneous put T and output T out torque values experienced by the transmission. With these measurements, the put and output power values are determed by 2π P =ω T = Ω T, (3.1a) 60 2π Ω P =ω T = T 60 Γ out out out out i, (3.1b) where Γ i is the put-to-output speed ratio at gear stage i. With these, the total transmission power loss under loaded conditions is determed simply as PT = P Pout. (3.2) 33

49 In order to separate load-dependent (mechanical) and load-dependent (sp) power losses, each loaded test was repeated with no output (reaction) load such that T = 0. In these unloaded tests, the bolts of the flexible couplg between the transmission and the output torque-meter were physically removed, disconnectg the output-side drive tra components Figure 2.2. In this case, T out = 0. Hence, if there were no power loss the transmission, no put torque would be needed (i.e. T = 0 ) for operatg the transmission at the desired speed of Ω. Yet, due to the load-dependent losses, the drive motor must apply a certa amount torque T 0 that is equal to the torque loss due to sp losses. The power loss this case is 2π Ps = P _ unloaded = Ω T. (3.3) 60 Comparg this P s value to the total loss from Eq. (3.2), the total (directly) measured mechanical power loss of the transmission is found as Pm = PT Ps. (3.4) Usg the total transmission power loss P T and the put power P, the overall transmission efficiency is determed as P η 1 T T =. (3.5) P Likewise, the mechanical efficiency is defed as 34

50 P η 1 m m =. (3.6) P It is also noted here that sp loss efficiency is not possible to defe sce there is no useful power transmitted by the gear tra. It is also safe to state that the efficiency values as defed by Eq. (3.5) and (3.6) should be rather low at lower T (and hence P ) values sce P m becomes relatively small compared to P s. Figure 3.1 shows the variation of the oil temperature θ and total power loss P T with time for an example test run at T = 50 Nm, Ω = 2000 rpm, and V = 2.0 liters. In this figure, the solid curve correspondg to the left hand side vertical axis shows the measured oil temperature θ () t, which varies from room temperature to nearly 90 C. The dashed curve correspondg to the right hand side vertical axis shows total power loss PT () t calculated from measured T and T out usg Eq. (3.2). It can be seen from this figure that P T = 2.6 kw at the begng of the test when θ= 30 C, and that value reduces exponentially to about 1.7 kw after one hour of testg. The oil temperature reaches nearly 90 C over this same time period. This figure demonstrates how P T values vary with time and oil temperature (viscosity) and also how they can be extracted at discrete temperatures from the contuous simultaneous time histories θ ( t) and P ( t ) collected. Here, the P T value at a given temperature θ a was obtaed by averagg the oil T P values of all measured data pots with the temperature range θ a ± 1 C so that T stantaneous variations P T are somewhat averaged out. 35

51 θ [ C] P T [ kw ] Temperature 0.5 Power Loss Time [m] Figure 3.1. A typical measured variation of the total power loss P T and transmission oil temperature θ a function of time. T = 50 Nm, Ω = 2000 rpm and V = 2.0 liters. oil 36

52 3.2 Transmission Efficiency Test Results Results of the experiments that were performed accordg to the test matrix of Table 2.1 are presented this section to demonstrate the measured fluence of the key operatg parameters Ω, θ, T and V oil on P T, P s and P m as well as η T and η m Influence of Input Speed Ω In Figure 3.2, measured values of P T, P sp, and P m are plotted as a function of gear stage (2 nd to 6 th ) when the transmission is operated at θ= 80 C, T = 50 Nm, V oil = 3.1 liters and Ω = 1, 000 4,000 rpm. In Figure 3.2(a), it is observed that the P T values crease with the gear stage, a trend that is more obvious at higher Ω values. For stance, the measured P T values at 2 nd gear stage are 0.42 and 2.26 kw for Ω = 1,000 and 4,000 rpm, respectively. The correspondg values of P T at the 6 th gear stage are 0.70 and 4.42 kw (67 and 96% higher than those at the 2 nd gear stage), dicatg the combed fluence of the gear stage and Ω. The correspondg P s values for the same conditions are shown Figure 3.2(b). Here, the variation of P s with the gear stage and Ω is more pronounced. For stance, at the 2 nd gear stage, P = 0.15 and 1.22 kw at Ω = 1,000 and 4,000 rpm, respectively. Meanwhile, at the s 6 th gear stage, P s = 0.47 and 4.20 kw at the same speed values, correspondg to about 67 and 95% of P T. This suggests that the sp losses constitute a much larger portion of P T durg operation at higher gear stages. Fally, the correspondg mechanical power 37

53 rpm 2000 rpm 3000 rpm 4000 rpm ( a ) P T [ kw ] P s [ kw ] rpm 2000 rpm 3000 rpm 4000 rpm ( b ) Gear Stage Figure 3.2. Measured (a) P T, (b) P s and (c) P m at T = 50 Nm, θ= 80 C and V = 3.1 liters. oil 38

54 P m [ kw ] rpm 2000 rpm 3000 rpm 4000 rpm () c Gear Stage Figure 3.2. Contued. Measured (a) P T, (b) P s and (c) P m at T = 50 Nm, θ= 80 C and V oil = 3.1 liters. 39

55 losses (difference of the values from Figures 3.2(a) and 3.2(b)) shown Figure 3.2(c) dimish with up-shift to higher gear states. In order to demonstrate the fluence of Ω on components of the transmission power loss as well as the relative amplitudes of P m and P s more directly, the data of Figure 3.2 is presented Figure 3.3 with Ω as the variable stead of the gear stage. In Figure 3.3(a), the measured variations of P T, P m and P s with Ω are shown when the transmission is at the 2 nd gear stage. It is noted here that P s is about 35% of P T at Ω = 1,000 rpm while it is nearly 54% at 4,000 rpm. The variation of P T with Ω follows the exponential relation c T = + Ω where the exponent c = 1.2 the second P a b gear stage. This figure dicates that, while P s becomes more domant at higher speed values, both P s and P m contribute to P T significantly. This behavior changes gradually as the transmission contues to up-shift, with P s takg a larger role. This is illustrated Figures 3.3(b-e) for the 3 rd to 6 th gear stages. In Figure 3.3(e), the same exponential relationship is evident with c = 1.3, and P s is nearly 95% of P T at 4,000 rpm. Likewise, the curve fit for P s reveals P = 0.45Ω, with the exponent value rather close to that of s P T. It is also observed these figures that P m remas somewhat flat with regardless of the gear stage considered. Figure 3.4 shows the overall η T and mechanical η m efficiency values correspondg to the power losses presented Figure 3.2. While η m values are above 94% at all gear stages and speed values as shown Figure 3.4(b), the same cannot be Ω

56 Total Power Loss Sp Loss Mechanical Loss ( a ) P [ kw ] ( b ) P [ kw ] Input Speed [rpm] Figure 3.3. Comparison of measured variation of P T, P s and P m with Ω at (a) 2 nd, (b) 3 rd, (c) 4 th, (d) 5 th and (e) 6 th gear stages. T = 50 Nm, θ= 80 C and V oil = 3.1 liters. 41

57 Total Power Loss Sp Loss Mechanical Loss () c 3.0 P [ kw ] ( d ) 3.0 P [ kw ] Input Speed [rpm] Figure 3.3. Contued. Comparison of measured variation of P T, P s and P m with at (a) 2 nd, (b) 3 rd, (c) 4 th, (d) 5 th and (e) 6 th gear stages. 50 and V oil = 3.1 liters. 42 T = Nm, θ= 80 Ω C

58 Total Power Loss Sp Loss Mechanical Loss () e P [ kw ] Input Speed [rpm] Figure 3.3. Contued. Comparison of measured variation of P T, P s and P m with at (a) 2 nd, (b) 3 rd, (c) 4 th, (d) 5 th and (e) 6 th gear stages. 50 and V oil = 3.1 liters. T = Nm, θ= 80 Ω C 43

59 1.00 ( a ) 0.95 η T rpm 2000 rpm 3000 rpm 4000 rpm 0.95 ( b ) 0.90 η m Gear Stage Figure 3.4. Measured (a) η T and (b) η m at T = 50 Nm, θ= 80 C and V oil = 3.1 liters. 44

60 said for the measured η T values Figure 3.4(a). For stance, for Ω = 4,000 rpm, η = 0.896, 0.881, 0.852, 0.806, and at 2 nd to 6 th gear stages, respectively. This T further emphasizes the creasg fluence of Ω at higher gear stages due to the correspondg creases P s values. The underlyg reason why P s creases with gear stage is simply the fact that the transfer shafts and the output shaft rotate at higher speed. As a result, the fal drive gear pairs rotate much faster under smaller loads, causg an crease P s values. Accordg to Figure 3.4(a), the efficiency of a manual transmission can be even lower than 0.8, dicatg that, some condition, more than one-fifth of the power can be lost while passg though the transmission Influence of Oil Temperature θ Figure 3.5 shows the fluence of the oil temperature on measured values of P T, P s, and P m at the 4th, 5 th and 6 th gear stages for a transmission operatg at T = 50 Nm, Ω = 2,000 rpm and V = 3.1 liters. Here, the power loss values are compared at three oil oil temperature values of θ = 30, 50 and 80 C. This figure dicates that the oil temperature fluences P s, and hence P T, significantly sce it impacts the viscosity of the oil. For stance, P s values Figure 3.5(b) are 44 to 49% lower at 80 C than those at 30 C. Increased oil temperature corresponds to reduced viscosity of the oil, and churng of low-viscosity oil by the gears and beargs causes lower sp power loss values. The changes P m with oil temperature are rather modest as shown Figure 3.5(c). 45

61 C 50C 80C ( a ) P T [ kw ] ( b ) 2.0 P s [ kw ] Gear Stage Figure 3.5. Measured fluence of θ on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 2,000 rpm and V = 3.1 liters. oil 46

62 C 50C 80C () c P m [ kw ] Gear Stage Figure 3.5. Contued. Measured fluence of θ on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 2,000 rpm and V = 3.1 liters. oil 47

63 Figure 3.6 compares the η T and η m values correspondg to the power loss values presented Figure 3.5. As an example, η T = 0.891, and at 80 C and 4 th, 5 th and 6 th gear stages. The correspondg values at 50 C are η T = 0.854, and The overall efficiency values at 30 C are even lower ( η T = 0.809, and at 4 th, 5 th and 6 th gear stages). This represents a fairly evenly distributed but significant drop efficiency with reduced temperature across the range studied. Also, observg the changes η m shown Figure 3.6(b) with temperature, it can be stated that the impact of θ on η T is mostly through its significant fluence on P s. Figure 3.7 shows the variation of measured P T, P s, and P m with θ for the same conditions of Figures 3.5 and 3.6 except Ω = 4, 000 rpm stead of 2000 rpm. While the qualitative behavior observed Figure 3.6 remas tact, the quantitative changes P T with θ are now more severe. In comparison to Figure 3.5(a), P T values are nearly doubled here maly due to the order of creases P s, as shown Figure 3.7(b). The correspondg η T and η m values for the same set of data are shown Figure 3.8(a) and (b), respectively. Here, the overall transmission efficiency at 30 C is only at 6 th gear stage and at 4 th gear stage. In summary, it is concluded from Figures 3.5 to 3.8 that operatg oil temperature (viscosity) is a ma parameter fluencg the transmission efficiency through its significant fluence on sp power losses. 48

64 C 50C 80C ( a) η T ( b) η m Gear Stage Figure 3.6. Measured fluence of θ on (a) η T and (b) η m at T = 50 Nm, Ω = 2,000 rpm and V = 3.1 liters. oil 49

65 C 50C 80C ( a) P T [ kw ] ( b) 4.0 P s [ kw ] Gear Stage Figure 3.7. Measured fluence of θ on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 4,000 rpm and V = 3.1 liters. oil 50

66 C 50C 80C ( c) P m [ kw ] Gear Stage Figure 3.7. Contued. Measured fluence of θ on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 4,000 rpm and V = 3.1 liters. oil 51

67 C 50C 80C ( a) η T ( b) 0.90 η m Gear Stage Figure 3.8. Measured fluence of θ on (a) η T and (b) η m at T = 50 Nm, Ω = 4,000 rpm and V = 3.1 liters. oil 52

68 3.2.3 Influence of Nomal Input Torque T In Figure 3.9, the measured fluence of the nomal put torque on P T and P m is shown at Ω = 2,000 rpm, θ= 80 C and V oil = 3.1 liters. The discrete T values were chosen as 50 and 100 Nm, with a set of tests done at no-load ( T = 0 )as well so that the correspondg P s values could also be quantified. In this case, the difference P T with T is attributed solely to the changes P m sce P s represents the loaddependent power loss, and therefore, remas the same at each torque level. In Figure 3.9(b), creasg T from 50 Nm to 100 Nm creased the P m values at least 50% regardless of the gear stage, hence makg the P m portion of P T more significant. For stance, at the 6 th gear range, P m 0.4 P T for 100 Nm while it is only P m 0.15 P T at 50 Nm. Likewise, at the 3 rd gear range, P m 0.6 P T for 100 Nm while it is only P m 0.45 P T at 50 Nm. This dicates that operation of the transmission at lower speeds under heavily loaded conditions would brg the mechanical power losses to the forefront of efforts to reduce transmission power losses. Figure 3.10, which shows the η T and η m values for the same case at T = 50 and 100 Nm, further supports this observation. Here, it is noted that measured η T and η m crease with creases T even though both P T and P m crease with it. This is because creases losses happen at a lower rate than the creases the put torque, resultg higher efficiency. The put power P is 53

69 doubled by doublg T from 50 to 100 Nm, while the resultg crease P T is only 30-48%. As a result, η T ranges from at the 3 rd gear stage to at the 6 th gear Nm 50 Nm 100 Nm ( a) P T [ kw ] Nm 100 Nm ( b) P m [ kw ] Gear Stage Figure 3.9. Measured fluence of T on (a) P T and (b) P m at θ= 80 C, Ω = 2,000 rpm and V oil = 3.1 liters. 54

70 Nm 100 Nm ( a) η T ( b) η m Gear Stage Figure Measured fluence of T rpm and V oil = 3.1 liters. η on (a) T and (b) η m at θ= 80 2,000 C, Ω = 55

71 stage when T = 100 Nm. This range is reduced to between at the 3 rd gear stage and at the 6 th gear stage when T = 50 Nm. This represents a sizable difference η T between these two torque values. In Figure 3.11, the measured fluence of T on P T and P m is shown at Ω = 4,000 rpm, while all other operatg conditions are kept the same as Figure 3.9 ( θ= 80 C and V oil = 3.1 liters). Here, creasg T from 50 to 100 Nm creased the P m values between 50% and 80% at the various gear stages, once aga makg the P m portion of P T somewhat more significant at the higher load value. For stance, at the 6th gear stage, P m 0.2 P T at 100 Nm while it is only P m 0.05 P T at 50 Nm. Likewise, at 56 the 3 rd gear stage, P m 0.5 P T at 100 Nm while it is only P m 0.3 P T at 50 Nm. Here P m represents a smaller proportion of P T all cases than it did at Ω = 2,000 rpm because P s is much larger at 4000 rpm. When compared directly to the 2000 rpm data from Figure 3.9, the P T values at 4000 rpm are consistently 2 to 3.5 times higher, mostly due to the crease P s. Likewise, Figure 3.12 shows the η T and η m values for the same case as Figure 3.11 at T = 50 and 100 Nm. Once aga, η T creases significantly with the crease T, this time even more dramatically than at Ω = 2,000 rpm. For T = 100 Nm, the efficiency ranges from at the 3 rd gear stage to at the 6 th gear stage. The correspondg values at 50 Nm are at the 3 rd gear stage to at the 6 th gear stage. The difference the 6 th gear stage is especially significant (a difference of 0.08).

72 Nm 50 Nm 100 Nm ( a) P T [ kw ] ( b) 4.0 P m [ kw ] Gear Stage Figure Measured fluence of T on (a) P T and (b) P m at θ= 80 C, Ω = 4,000 rpm and V oil = 3.1 liters. 57

73 Nm 100 Nm ( a) 0.90 η T ( b) η m Gear Stage Figure Measured fluence of T on (a) η T and (b) η m at θ= 80 C, Ω = 4,000 rpm and V oil = 3.1 liters. 58

74 3.2.4 Influence of Oil Volume V oil The amount of oil put the transmission defes the static oil level with the transmission. It is a parameter that must be defed by product engeers based on applications hand. While it has been well known that creasg the oil levels leads to higher power losses due to creased oil churng activity, it also provides additional drag and dampg to help avoid transmission gear rattle problems. It also eases the burden on lubrication system design and lubricant performance requirements, as there is plenty of oil to carry out these tasks. As energy concerns become more proment, these advantages can be brushed aside as long as the efficiency improvements with reduced oil level are tangible. In this section, sets of experimental data will be presented an attempt to quantify such efficiency improvements as a function of oil volume. Based on field experience with the example transmission, three oil levels shown Figure 2.9 will be achieved by fillg the transmission with oil quantities of V oil = 3.1 (high), 2.0 (medium) and 1.2 liters (low). Figure 3.13 shows the fluence of V oil on measured values of P T, P s, and P m at the 4 th to 6 th gear stages when the transmission is operatg at Ω = 2,000 rpm, T = 50 Nm and θ= 30 C. As expected, the results show very little difference P m, while P s is creased significantly with creased V oil values. For stance, at the 6 th gear stage, P = 2.43, 2.06 and 1.44 kw for V = 3.1, 2.0 and 1.2 liters, respectively. This s oil represents a 15% decrease from 3.1 to 2.0 liters and a 30% decrease from 2.0 to 1.2 liters 59

75 (a 41% decrease P s when V oil is reduced from 3.1 to 1.2 liters). Differences are somewhat less at lower gear ranges sce the rotational speed of the differential gears is L 2.0 L 3.1 L ( a) P T [ kw ] ( b) 2.0 P s [ kw ] Gear Stage Figure Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 2,000 rpm and θ= 30 C 60

76 L 2.0 L 3.1 L ( c) P m [ kw ] Gear Stage Figure Contued. Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 2,000 rpm and θ= 30 C 61

77 lower. At the 4 th gear stage, P s = 1.47, 1.34 and 1.13 kw for oil volumes of 3.1, 2.0 and 1.2 liters. This represents a 9% decrease from 3.1 to 2.0 liters and a 16% decrease from 2.0 to 1.2 liters. In Figure 3.14, measured changes η T and η m values with V oil are shown for the same conditions of Ω = 2, 000 rpm, T = 50 Nm and θ= 30 C. This figure shows a fairly steady decrease η T with creasg V oil, while η m remas rather flat. A larger difference is observed the 6 th gear stage than the 4 th gear stage. The efficiencies 4 th gear are η T = 0.809, and for V oil = 3.1, 2.0 and 1.2 liters respectively, which shows a 5% spread solely due to the change V oil. This represents a 2.7% crease efficiency from 3.1 to 2.0 liters and a 2.2% crease from 2.0 to 1.2 liters. The overall efficiencies 6 th gear are 0.758, and at 3.1, 2.0 and 1.2 liters of oil, correspondg to a 3.3% crease η T from 3.1 to 2.0 liters and a 4.5% crease η T from 2.0 to 1.2 liters. Figures 3.15 and 3.16 provide the same comparison as Figures 3.13 and 3.14, now at an elevated oil temperature of θ= 80 C. The general trends rema the same as earlier, except the magnitudes of P s and P T are more modest, as are the changes them with V oil. This is a direct result of reduction churng losses due to reduced viscosity [2, 3]. At the 6 th gear stage, P s = 1.242, and kw for V oil = 3.1, 2.0 and 1.2 liters. This represents a 21.4% decrease from 3.1 to 2.0 liters and a 21.4% decrease from 2.0 to 1.2 liters. Once aga, changg the oil level has only a very small effect on P m and η m. 62

78 L 2.0 L 3.1 L ( a) 0.90 η T ( b) 0.90 η m Gear Stage Figure Measured fluence of V oil on (a) η T and (b) η m at T = 50 Nm, Ω = 2,000 rpm and θ= 30 C. 63

79 L 2.0 L 3.1 L ( a) P T [ kw ] ( b) P s [ kw ] Gear Stage Figure Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 2,000 rpm and θ= 80 C 64

80 L 2.0 L 3.1 L ( c) P m [ kw ] Gear Stage Figure Contued. Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 2,000 rpm and θ= 80 C 65

81 ( a) 0.90 η T L 2.0 L 3.1 L 0.95 ( b) η m Gear Stage Figure Measured fluence of V oil on (a) η T and (b) η m at T = 50 Nm, Ω = 2,000 rpm and θ= 80 C 66

82 Figures 3.17 and 3.18 provide another comparison at a lower temperature ( θ= 30 C) and higher put speed ( Ω = 4,000 rpm). Here, overall impact of V oil is more significant than the previous two comparisons at lower speeds and/or higher temperatures. For stance, at the 6 th gear stage (30 C and 4000 rpm), P s = 5.29, 4.05 and 3.52 kw for V oil = 3.1, 2.0 and 1.2 liters. This represents about 200 to 250% higher P values compared to those Figure 3.13 at θ= 30 C and Ω = 2,000 rpm. The s difference between Figure 3.15 ( θ= 80 C and Ω = 2,000 rpm) and Figure 3.17 ( θ= 30 C and Ω = 4,000 rpm) is even more strikg, as it amounts to 4 to 5 times larger losses for the latter. For the purpose of completeness, one last comparison of the fluence of V oil on power losses and efficiency is presented Figures 3.19 and 3.20, respectively, for θ=80 C and Ω = 4,000 rpm. The same trends observed before exist consistently here as well, providg a more complete picture of the combed fluence of three key parameters: Ω, θ and V oil. As a general trend, creases Ω and V oil as well as a decrease θ are all detrimental to the overall efficiency of the transmission. An example direct comparison is provided Figure 3.21 to illustrate this combed fluence. This is shown by choosg P T and η T values the 6 th gear stage for all four cases presented Figures 3.13 to It is observed from this figure that a 4.5 kw 67

83 L 2.0 L 3.1 L ( a) P T [ kw ] ( b) 4.0 P s [ kw ] Gear Stage Figure Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 4,000 rpm and θ= 30 C 68

84 L 2.0 L 3.1 L ( c) P m [ kw ] Gear Stage Figure Contued. Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 4,000 rpm and θ= 30 C 69

85 L 2.0 L 3.1 L ( a) η T ( b) 0.90 η m Gear Stage Figure Measured fluence of V oil on (a) η T and (b) η m at T = 50 Nm, Ω = 4,000 rpm and θ= 30 C 70

86 L 2.0 L 3.1 L ( a) P T [ kw ] ( b) P s [ kw ] Gear Stage Figure Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 4,000 rpm and θ= 80 C. 71

87 L 2.0 L 3.1 L ( c) P m [ kw ] Gear Stage Figure Contued. Measured fluence of V oil on (a) P T, (b) P s and (c) P m at T = 50 Nm, Ω = 4,000 rpm and θ= 80 C. 72

88 L 2.0 L 3.1 L ( a) η T ( b) 0.90 η m Gear Stage Figure Measured fluence of V oil on (a) η T and (b) η m at T = 50 Nm, Ω = 4,000 rpm and θ= 80 C. 73

89 6 5 4 ( a) 3.1 Liters 2.0 Liters 1.2 Liters P T [ kw ] ( b) 0.9 η T rpm 30 C 4000 rpm 80 C 2000 rpm 30 C 2000 rpm 80 C Figure Combed fluence of V oil, Ω and θ on (a) P T and (b) η T at 6 th gear range and T = 50 Nm. 74

90 spread (from 0.8 to 5.3 kw) P T and an 18% spread (0.74 to 0.92) η T exist, which is very significant. This suggests that these three are the major parameters dictatg the efficiency of the manual transmission. 75

91 CHAPTER 4 SUMMARY AND CONCLUSIONS 4.1 Summary In this study, a test methodology was developed to measure power losses of an automotive manual transmission under loaded and unloaded conditions. A set of special fixtures was designed and procured to operate the transmission with a dynamometer facility. A pair of precision torque-meters were acquired and implemented with the fixtures to measure the put and output torque (and hence power) values. The fidelity of the measured quantities and repeatability of the measured losses were demonstrated. An extensive test matrix that cludes various combations of four variables, speed, torque, oil temperature and oil volume, was defed and executed to quantify their fluence on total, mechanical and sp power losses. At the end, several observations were made based on the results of the tests. 76

92 4.2 Conclusions Ω, V oil Specifically: The analysis of the power loss data revealed that all four operatg parameters,, θ and T, fluence the power losses from the transmission significantly. P s significantly creases with creases Ω and V oil. The crease P s with Ω is exponential. Likewise, creased oil volume causes creased oil churng activity that results elevated P s values. While the range of 77 Ω is not negotiable, reducg the oil volume with the transmission to a mimum amount that still meets the durability requirements can be identified as a way of creasg transmission efficiency. Influence of Ω on P s is more pronounced at higher gear stages sce the transfer and output gear sets rotate at higher speeds. P s creases with reduced oil temperature θ. The sp losses at cold-start (room temperature) conditions are significantly larger than those at typical steady-state operatg temperatures. This is maly due to the crease oil churng losses with creased viscosity at lower operatg temperature ranges. As expected, T primarily fluences the P m portion of P T, as creased amounts of nomal load transmitted correspond to creases P m. Losses the form of P m represent a larger percentage of P T at lower gear ranges, while they become secondary at higher speed and higher gear ranges of

93 operation. This suggests that the efforts to improve the transmission efficiency at high speed and lower load conditions (as highway drivg conditions) must focus on reduction of sp power losses rather than friction-duced mechanical losses. 4.3 Recommendations for Future Work The experimental database generated this study exhibited clear trends terms of the fluence of various operatg parameters on transmission power losses. While these trends can be described heuristically, it is not possible to ppot the underlyg mechanisms leadg to such trends. In order to brg a complete understandg to the measured power loss behavior of manual transmission efficiency, this experimental work must be complemented by a theoretical study. While this experimental study provided a thorough examation of the parameters studied, several expansions to the database are warranted. Among them, the nomal torque T range should be creased. The torque range that was used this study was limited by the capabilities of the dynamometer facility used this study. Likewise, the range of oil volume can also be expanded beyond the range used production to obta a more complete picture of this trend. There is also room for expandg the range of Ω beyond 4,000 rpm sce modern ternal combustion enges reach speeds beyond that. In terms of the measurement system, potential improvements can be implemented by usg torque-meters that have higher sensitivity and accuracy. The torque-meters used 78

94 this study had a resolution of 0.25% of the maximum torque capacity. This corresponds to 0.25 Nm for the put torque-meter while it is 3.75 Nm for the output torque-meter. While significant trends were possible to quantify with such resolution, secondary effects, especially concerng the mechanical losses, often fall below this resolution. 79

95 BIBLIOGRAPHY [1] Seetharaman, S., Kahraman, A., Bednarek, G. and Rosander, P., 2008, A Model to Predict Mechanical Power Losses of Manual Transmissions, Automobiltechnische Zeitschrift, April 2008, Issue 4, pp [2] Seetharaman, S., Kahraman, A., 2008, Load Independent Sp Power Losses of a Spur Gear Pair: Model Formulation, ( review) Journal of Tribology. [3] Seetharaman, S., Kahraman, A., Moorhead, M. D., and Petri-Johnson, T. T., 2008, Load Independent Sp Power Losses of a Spur Gear Pair: Experiments and Model Validation, ( review) Journal of Tribology. [4] Mart, K. F., 1978, A Review of Friction Predictions Gear Teeth, Wear, 49, pp [5] Yada, T., 1997, Review of Gear Efficiency Equation and Force Treatment, JSME Int. J., Ser. C, 40, pp [6] Li, Y., and Seireg, A. A., Predictg the Coefficient of Friction Slidg-Rollg Contacts, Tribology Conference, K18. [7] Hegartner, P, and Mba, D., 2003, Determg Power Losses the Helical Gear Mesh, Gear Technology, September/October 2005, pp [8] Changenet, C., Oviedo-Marlot, X., and Velex, P., 2006, Power Loss Predictions Geared Transmissions Usg Thermal Networks-Applications to a Six-Speed Manual Gearbox, Transactions of the ASME, Vol. 128, pp [9] Handschuh, R. F., Kilma, C. J., 2003, Efficiency of High-Speed Helical Gear Tras, 59 th Annual Forum and Technology Display sponsored by the American Helicopter Society, Phoenix, AZ. [10] van Dongen, L. A. M., 1982, Efficiency Characteristics of Manual and Automatic Passenger Car Transaxles, SAE Technical Paper Series, , Warrendale, PA: Society of Automotive Engeers. [11] Barzaghi, C., Berti, F., and Gommelli, C., 1995, "Development of a Bench Test Procedure for Assessg the Effect of Lubricants on Car Manual Transmission Efficiency," S.A.E. Transaction, 104, 4, 668. [12] Greenbaum, J. J., Kluger, M. A., and Westmoreland, B. E., 1994, Manual Transmission Efficiency Trends and Characteristics, SAE Technical Paper Series, , Warrendale, PA: Society of Automotive Engeers. 80

96 [13] Kluger, M. A., Greenbaum, J. J., Mairet, D. R., 1995, Proposed Efficiency Guideles for Manual Transmissions for the Year 2000, SAE Technical Paper Series, , Warrendale, PA: Society of Automotive Engeers. [14] Kluger, M. A., Long, D. M., 1999, An Overview of Current Automatic, Manual and Contuously Variable Transmission Efficiencies and Their Projected Future Improvements, S.A.E. Transaction, 108, pp

97 APPENDIX A DESIGN SPECIFICATIONS OF MAJOR MACHINE ELEMENTS 82

98 Figure A.1. Input and output torque-meter specifications (source: 83

99 Figure A.2. Input and output torque-meter dimensions (source: 84

100 Figure A.3. Flexible couplg specifications (source: 85

101 Figure A.4. Flexible couplg dimensions (source: 86