Improving the Efficiency and Lowering the Operating and Manufacturing Costs by Suitable Power Distribution of Medium Speed Two Stage Planetary Gearboxes for Next Generation Wind Turbines ATTILA CSOBÁN Department of Machine and Product Design Budapest University of Technology and Economics Műegyetem quay 3. HUNGARY csoban.attila@gt3.bme.hu Abstract: - The main requirements having to be met by the gear drives are high reliability and load carrying capacity, long lifespan, high efficiency, easy manufacture and assembly, as well as low costs. These requirements can be met by some two stage compound planetary gears. Combining the elements of two stage planetary gears it is possible to create more than 140 constructions. Comparing the efficiency, volume and operating costs it can be seen that for next generation medium speed turbine drive train in the middle gear ratio range the simple two stage planetary gear type3 (Fig. 3.) is the best construction form all the 140 concepts. The manufacturing cost can be reduced by using the maximum number of planetary gear with smaller bearings. To lowering the operating costs, reach higher efficiency and reducing mass more power distribution and torque split is needed. I worked in the field of design and I invented a new type of, which has a higher power distribution and less volume comparing to common types of gears. Revolutionary new drive designs based on patented (P1400594) new concepts and characterized by higher efficiency, smaller form factor and smaller weight compared to traditional drives. Key-Words: - Wind turbine, downsizing, planetary gear, power distribution, bearing cost, medium speed es 1 Introduction Looking the wind turbine market there is an exponential growth in the power generated by wind turbines. In year 2000 the cumulative installed wind power in the world was about 5.000 [MW/Year] but till 2010 it reached almost 40.000 [MW/Year]. Offshore market share forecast to increase to 9.6 [%] of total installed capacity by 2015 [2]. Wind farms generate power with relatively good efficiency and with relatively low costs comparing with other alternative renewable energy sources but there are very strict requirements. Seeing the trends a monotonous growth in the size of installed turbines can be observed. Forecasts indicate the largest market share will be for turbines generate more than 3 [MW] power. The highest mechanical power of existing offshore turbines is 5 [MW] which is much bigger than the global average: 1,65 [MW]. Current development is focusing on turbines having more than 165 [m] rotor diameter and generating 7.5 [MW] or more power. If the trend will not change, the size of next generation turbines will reach 10 [MW] or more power having a rotor diameter of 200 [m]. To win the challenge and to define the ways of drivetrain innovation it is necessary to install bigger turbines and lower all costs of the drive system. There can be three different ways of the transmission design. The most dominant ways of the wind turbine transmission innovations are high speed modular es having a gear ratio approximately 1:120, direct drive drivetrains and relatively new medium speed drivetrains with ratios from 1:10 up to 1:40. Using the conventional high speed transmissions it is possible to transmit 1-5 [MW] power just the same as with direct drive, considering the limitation in weight and size. Taking into account the trends the medium speed geared transmissions can be used [1]-[9] for turbines planned to transmit more than 10 [MW] power. The main parts of these types of are the water cooled permanent magnet generator, the compound planetary gear and the main bearing integrated into one unit. The main requirements having to be met by the gear drives are high reliability and load carrying capacity, long lifespan, high efficiency, easy manufacture and assembly, as well as low costs. ISBN: 978-1-61804-314-6 162
These requirements can be met by some two stage compound planetary gears. Combining the elements of two stage planetary gears it is possible to create more than 140 constructions. Investigating all the 140 constructions and taking into account the needed gear ratio range and high efficiency, it can be proved that only 6 constructions are suitable to meet the requirements mentioned above. Beside the high load carrying capacity and the low noise emission of the es it is more and more important to reach their highest efficiency in order to lower the power requirement and the heat development. Although the efficiency of a modern planetary gear drives is usually over 95 % at nominal load [10]-[20] and optimal parameters such as inner gear ratios, tooth profiles, lubrication systems and lubricants. In this article the 6 compound planetary were investigated and compared to find the best construction for next generation wind turbines. Fig.1. Planetary gear type 1. First stage ring gear is the input shaft, output on second stage sun gear. Fig.2. Planetary gear type 2. First stage carrier is the input shaft, output on second stage sun gear. 2 Medium Speed Two Stage Planetary Gear Constructions There are some types of planetary gears which are able to result high gear ratio, but their power flow can be unbeneficial, because a large part of the rolling power can circulate inside the planetary decreasing the efficiency. In the simple planetary gears there is no idle power circulation. Heavy-duty planetary drives consisting of simple planetary gears are planned to transmit megawatts or even more power, while they must be compact and efficient. These requirements can be solved with two stage compound planetary gears. Combining the elements of two stage planetary gears it is possible to create more than 140 constructions. Investigate all the 140 constructions and take into account the needed gear ratio range and high efficiency, only 6 constructions are able to solve the requirements. These types can be seen in Figure 1 6. Fig. 3. Planetary gear type 3. First stage carrier is the input shaft, output on second stage sun gear. Figures 1-3 show two stage planetary gears consisting of two simple planetary gears KB coupled to each other. In this case varying the inner gear ratio (the ratio of tooth number of ring gear and sun gear) of the simple planetary gear units KB (Figure 1-3.) the performance of the two stage planetary gear can be changed and tailored to the requirements. There are other special types of planetary gears also consisting of simple KB units, which can divide the applied power between the planetary stages thereby increasing the specific load carrying capacity of the two stage planetary drives (Figures 4-6). Creating right connections between the elements of the two stages in these differential planetary gears there is no idle power circulation. The gear ratio, efficiency and the estimated costs of these types of two stage and differential planetary gears are investigated in the following, structure of which can be seen in Figure 1-6. ISBN: 978-1-61804-314-6 163
drive. The efficiency of the gears were calculated with a simple rolling efficiency takes into account the teeth and bearing friction losses in each stage of the. Fig.4. Planetary gear type 4. First stage carrier and second stage ring gear is the input shaft, output on first stage sun gear. Fig.5. Planetary gear type 5. First stage and second stage ring gear is the input shaft, output on first stage sun gear. Fig. 6. Planetary gear type 6. First step ring gear is the input shaft, output on second stage sun gear. 3 Problem Solution The performance of a planetary gear drive depends on its kinematics, its inner gear ratios and the connections between the planetary stages. Only detailed calculations can reveal the behavior of planetary gears and show the best solution for a given application. For comparing the gear drives the gear ratio and gear efficiency and also the mass of planetary gears (Figure 1-6) were calculated. For example for gear type 3 the gear ratio were calculated with the following equation (Fig. 3): 1 itype 3 (1) 1 ib " 1 ib ' The efficiency of planetary gear type 3: 1 ib' 1 ib" itype 3 (2) i b' i b" 1 1 g ' g" To determine the sizes of the es the center distances of the mating gears had to be known which was calculated using the well-known equations taking into account the Hertzian contact stresses between the teeth. The rolling efficiency values were assumed to be 97 [%] during the calculations taking into account the tooth, the bearing and the oil churning losses. When the geometry of the main parts like sun, planet and ring gears were known, the mass of each stages were calculated using a simple model shown in Fig. 7. 2 2 2 2 m234 k b r2 N r3 0,5625 r4 a w (3) The mass of the planetary es: m (4) m 234kj j1 l 2.1 Efficiency of Planetary Gears The sources of energy losses of a planetary are friction loss between the mating teeth, friction loss in the bearings, friction loss at the seals, energy losses owing to lubricant churning, energy loss of air-drag. The main source of energy loss is the tooth friction of gears depending on the arrangements of the meshing gears and the power flow inside the planetary gear drives. The bearing friction loss is the second biggest power loss generated inside the gear Fig.7. The main parts of a simple planetary gear unit and the simplified model (2 sun gear, 3 planet gears, k carrier, 4 ring gear). In the equations the following notations are used: ib - is the ratio of the numbers of teeth of the ring gear and the sun gear at the second stage, ib - is the ISBN: 978-1-61804-314-6 164
ratio of the numbers of teeth of the ring gear and the sun gear at the first stage, g- is the product of tooth efficiencies in a simple planetary gear stage, i - is the gear ratio, 2- sun gear, k- planetary carrier, 3- planet gear, 4- ring gear, l- the number of the stages, N- the number of the planet gears, r- the extended pitch circle of gears, aw- center distance, b width of gear, the density of steel, m mass of gear drive, d m bearing average diameter, M relative mass reduction, C relative cost reduction, r relative ring gear radius reduction. Behaviors of these types of two-stage and differential planetary gears were investigated and compared using the derived equations and following the train of a systematical thought. The first step was to choose the inner gear ratios of every stage and to combine them creating as many planetary gear ratios as possible. Using the equations developed the efficiency and the mass of every gear can be calculated. If we take into account that there is close relationship between the cost of manufacturing and mass of the used material and the machined areas it is important to see weather using the maximal power distribution in each stage worth it or not. The maximal power distribution can be reached by using the maximum number of planet gears in each stage. There can be another advantage of using the maximum number of planet gears in a planetary. Thanks to the power distribution size reduction can be made and not only the diameter of the gears but the size of the bearings can be smaller. The bearing cost [21] is also reduced by downsizing the.the cost of radial cylindrical roller bearing as a function of size is shown in Fig. 8. a relative mass reduction can be determined with the following equation: m M N Nmax mn 3 m N 3 (5) In practice the manufacturing cost of a gear or shaft is at least: 31,86 [ /kg]. If take into account that with higher power distribution the number of main parts are growing the difference in prices will lower but almost not equal. The sum of relative price reduction of using maximum power distribution was calculated with the following equation: C C N Nmax C N 3 C N 3 (6) Also the relative radius difference was calculated. The relative radius difference means the ring gear radius difference between the stages divided by the radius of the first stage ring gear. This parameter was calculated with the following equation: r 4 r 4" r r 4' 4' (7) The result for planetary gear type 3 is presented in diagrams (Fig 9-11). Comparing the calculated values of efficiency and relative cost reduction in the gear ratio range between 10:1 and 40:1 the optimal construction can be selected. The calculations were performed for all the six types of planetary gears planned to transmit a power of 750 kw at a driving speed of 10 rpm. ib'=1,5 ib'=2 ib'=2,5 ib'=3 ib'=3,5 ib'=4 ib'=4,5 ib'=5 ib'=5,5 ib'=6 ib'=6,5 ib'=7 ib'=7,5 ib'=8 ib'=8,5 ib'=9 ib'=9,5 ib'=10 96,0% 95,8% [%] 95,6% 95,4% Fig. 8. Cost of cylindrical roller bearing as a function of the inner radius. Comparing the mass of a construction have only three planet gears in each stage to the mass when the maximum number of planet gears are used 95,2% 95,0% 0,025 0,045 0,065 0,085 i [-] Fig.9. The efficiency of planetary gear type3 as a function of gear ratio. ISBN: 978-1-61804-314-6 165
ib'=1,5 ib'=2 ib'=2,5 ib'=3 ib'=3,5 ib'=4 ib'=4,5 ib'=5 ib'=5,5 ib'=6 ib'=6,5 ib'=7 ib'=7,5 ib'=8 ib'=8,5 ib'=9 ib'=9,5 ib'=10 Comparing the weight of the gears the mass of the gears were divided by the mass of gear type1. 0,025 0,050 0,075 i [-] 0,100 0% -5% -10% 160 120 M/Mgear [%] type4 type5 type1 type6 type2 type3-15% C[%] -20% -25% -30% 80 40-35% -40% Fig.10. The total cost reduction of planetary gear type3 as a function of gear ratio. ib'=1,5 ib'=2 ib'=2,5 ib'=3 ib'=3,5 ib'=4 ib'=4,5 ib'=5 ib'=5,5 ib'=6 ib'=6,5 ib'=7 ib'=7,5 ib'=8 ib'=8,5 ib'=9 ib'=9,5 ib'=10 100% 0 i=10 i=20 i=30 i=40 Fig.13. Relative mass of the es. The biggest ring gear radius of the divided by the biggest ring gear radius of gear type1 shows a relative size difference between the gears. r4 [%] 80% 60% 40% 160 120 Rmax/Rgear [%] type4 type5 type1 type6 type2 type3 20% 80 0% 0,03 0,04 0,06 0,07 0,09 0,10 40 i [-] Fig.11. The relative radius difference between the radius of ring gears. Gear type3. Comparing the efficiencies of the gears by discrete gear ratios can be seen in fig.12-14. 0 i=10 i=20 i=30 i=40 Fig.14. Relative size of the es. 96 95 94 93 Efficiency [%] type4 type5 type1 type6 type2 type3 i=10 i=20 i=30 i=40 Fig.12. Efficiencies of es. 4 Conclusion From the calculated results it can be seen that for next generation turbine drive train in the middle gear ratio range the simple two stage planetary gear type3 is better although there is no power distribution between the stages. The power flow and the construction is simple, the input element is the planetary carrier of the first stage and the output of the first stage is the sun gear which is connected with the planetary carrier of the second stage. The output element is the sun gear of the second stage. The manufacturing cost can be reduced by using the maximum number of planetary gear with smaller bearings. These types of es are well known ISBN: 978-1-61804-314-6 166
and with optimal tooth geometry a relatively good efficiency can be reach with relatively low manufacturing and operating costs. From 140 gear construction in this ratio range type3 is the best concept for medium speed turbines. To lowering the operating costs, reach higher efficiency and reducing mass more power distribution and torque split is needed. I worked in the field of design and I invented a new type of, which has a higher power distribution and less volume comparing to common types of gears. Revolutionary new drive designs based on patented (P1400594) new concepts and characterized by higher efficiency, smaller form factor and smaller weight compared to traditional drives. The 50 patented concepts guarantee that an optimum implemented shape is available for every application, whether with fixed or variable transmission. Depending on the application area, a lot of costs can be saved thanks to the higher efficiency and lower mass. Fig.17. Mass reduction with the new gear drives. Fig.15. The 3D model and the first prototype of the patented. Comparing the designed new construction with some common a lot of cost, mass and energy can be saved. Sale price of surplus energy generated by higher efficiency, calculated for 7200 hours of operation per year (unit price: 0,08 [USD/kWh]) is represented in Fig.16. Fig.18. The efficiency of the new gear drive is higher than the efficiency of common types of planetary gears. Fig.16. Cost saving with better efficiency. References: [1] Fingersh, L., M. Hand, and A. Laxson. Wind Turbine Design Cost and Scaling Model. Technical report, Golden: National Renewable Energy Laboratory, 2006. [2] John Coultate, Wind Turbine Drivetrain Technology and Cost Drivers, Technical report, www.romaxtech.com [3] Devendra Singh, Dr. Mohd. Suhaib: Kinematic Considerations in Gear Drives A Review, International Journal of Innovative Research in Science, Engineering and Technology (An ISO 3297: 2007 Certified Organization), Vol. 3, Issue 1, January ISBN: 978-1-61804-314-6 167
2014 [4] Hau, E. Wind Turbines: Fundamentals, Technologies, Application, Economics. 2nd Edition. New York: Springer, 2006. [5] Mueller, M., and A. McDonald. A lightweight low speed permanent magnet electrical generator for directdrive wind turbines. Proceedings of the 2008 European Wind Energy Conference & Exhibition. [6] Musial, W., S. Butterfield, and B. McNiff. Improving Wind Turbine Gearbox Reliability. Conference paper, Golden: National Renewable Energy Laboratory, 2007. [7] Poore, R., and T. Lettenmaier. Alternative Design Study Report: WindPACT Advanced Wind Turbine Drive Train Designs. Golden: National Renewable Energy Laboratory, 2003. [8] Ragheb, A., and M. Ragheb. Wind Turbine Gearbox Technologies. Proceedings of the 1st International Nuclear and Renewable Energy Conference (INREC10). Amman, Jordan, 2010. 8. [9] S.El Ahmadi, A. Benaboud, M.El Gameh, A.Echchelh, A.Chaouch: The Cost and the Cost-Effectiveness of Renewable Energy in Rural Areas, International Journal of Research Studies in Science, Engineering and Technology, Volume 1, Issue 8, November 2014, PP 30-37, ISSN 2349-4751 (Print) & ISSN 2349-476X (Online) [10] Csobán A., Kozma M.: Investigation the Energy Losses Generated by the Oil Churning, the Tooth and Bearing Friction in a Wolfrom Planetary Gear.TRIBOLOGIE UND SCHMIERUNGSTECHNIK 3/10:(57) pp. 32-35. (2010) [11] Csobán A.: The Bearing Friction of Compaund Planetary Gears int he Early Stage of Design for Cost Saving and Efficiency, TRIBOLOGY LUBRICANTS AND LUBRICATION, Edited by Chang-Hung Kuo, Open Access, INTECH, Book Chapter 4, pp. 119-138, ISBN 978-953-307-371-2 [12] Csobán A., Kozma M.: Investigation of the Bearing Friction Losses of Heavy- Duty Planetary Gears. ÖTG Symposium 2008 ISBN 978-3-901657-30-6, Wr Neustadt, Austria, 20 Nov 2008, pp. 259-266. [13] Dirk Strasser: Einfluss des Zahnflanken- und Zahnkopfspieles auf die Leerlaufverlustleistung von Zahnradgetrieben, Dissertation zur Erlangung des Grades Doktor- Ingenieur, Fakultät für Maschinenbau, Ruhr-Universität Bochum, 2005 [14] Csobán A., Kozma M.: Influence of the Oil Churning, the Bearing and the Tooth Friction Losses on the Efficiency of Planetary Gears. STROJNISKI VESTNIK-JOURNAL OF MECHANICAL ENGINEERING 4:(56) pp. 245-252. IF: 0.533* [15] Burton, T., Sharpe, D, Jenkins, N, Bossany, E. (2004). Wind Energy Handbook (3rd Ed.). John Wiley & Sons Ltd., ISBN: 0-471-48997-2, West Sussex, England. [16] Musial, W. Butterfield, S., McNiff, B. (2007). Improving Wind Turbine Gearbox Reliability, Proceedings of the 2007 European Wind Energy Conference, NREL: CP-500-41548, Milan, Italy, May 2007. [17] Mangliardi, L, Mantriota, G. (1994). Automatically Regulated C.V.T. in Wind Power Systems. Renewable Energy, Vol. 4, No. 3, (1994), pp. 299-310, 0960-1481(93)E0004-B. [18] Mangliardi, L., Mantriota, G. (1996). Dynamic Behaviour of Wind Power Systems Equipped with Automatically Regulated Continuously Variable Transmission. Renewable Energy, Vol. 7, No. 2, (1996), pp. 185-203, 0960-1481(95)00125-5. [19] Mikhail, A.S., Hahlbeck, E.C. Distributed Power Train (DGD) [20] Ragheb A., Ragheb, M. (2010). Wind Turbine Gearbox Technologies, Proceedings of the 1st International Nuclear and Renewable Energy Conference (INREC 10), ISBN: 978-1- 4244-5213-2, Amman, Jordan, March 2010. [21] SKF Price List, SKF ERP 2014_tcm_12-153040 pricelist 2014. ISBN: 978-1-61804-314-6 168