Ramanaiah, G. V.

English

Periodica Polytechnica (Budapest University of Technology and Economics)

ON THE ACCURACY AND RAPIDITY OF EVALUATION OF ONE·NODE NATURAL FREQUENCY AND ANGULAR AMPLITUDES OF UNDAMPED MULTICYLINDER ENGINE SYSTEMS By G. V. RAl\IANAIAH Department of Gas Engines and Automobiles. Technical University. Budapest (Received April 2. 1968) Presented by Prof. Dr. A. JL:REK 1. Introduction Several methods have been developed to find the one~node frequencies of undamped engine systems. Some of them are yery accurate but more time eonsuming. Other methods have improved quickness of evaluation but mostly approximate. It is always the aim of an engineer to combine both these diverse aspects and thus achieve rapidity as well as accuracy of assessment. True that some of these methods are not for the benefit of a newcomer but for the use of an experienced engineer who is 'very much pressed for time. The relative merits and demerits of different methods are discussed in the present paper and suggestions are given how to improve the accuracy of evaluation with reasonable rapidity. The different procedures to evaluate the one-node natural frequency of a given system broadly fall into two categories. (i) Methods which require the frequency equation and (ii) Those which do not. In the former group it is required just to solve an algebraic equation of higher order for its roots and this goes into the area of algebra. Forming the higher order algebraic equation itself is an interesting problem for an engineer also. One of the methods to form them from a determinant of specific form which occurs in the case of a multicylinder inline engine is given by CROSSLEY and GER;\fEN [1]. 2. The polynomial equation Considering an n-mass system sho ... ~-n in Fig. 1 the equations of motion could be 'written as =0 J D I k (0 O)! 1. 10 0) - 0 20~ T 1 02 - ClI T n2\(7::! - 03 -(1) (2) 368 G. V. RAMANAIAH (3) etc. until I n en + k"-1 (en - en-I) =0 (4) where kl' k2' kJ' . .. , kn - 1 are the spring constants of the connecting shafts, J 1, J 2, J3, ... , In the polar moments of inertia, el' e 2, e 3, •.. , en the angular amplitudes of different masses. Fig, 1 If the system is assumed to be vibrating simple harmonically 'with a natural frequency (jJ then e= - (jJ2e Substituting 5 in equations 1, 2, 3, and 4 (k - J (jJ2) eke 1 1 1- 1 2 =0 - kl e 1 + (kl + k2 - J2(jJ2) e 2 - k2e 3 = 0 - k2 e 2 + (k2 + k3 - J,3 (jJ2) e 3 - k3 e.1 = 0 etc. until =0 Equations 6 to 9 can be "\VTitten in the form of a determinant "where L1 all = k1 - 11 (1)2 a21 = - kl a32 = - k2 e 1 e 2 e 3 ... e n all a12 0 ... 0 a21 a22 a23 •.• 0 0 a32 a33 0 =0 etc. until 0 0 0 ann U12 = - kl a22 = kl + k2 - J2(jJ2 a23 = - kz (5) (6) (7) (8) (9) (10) EVALUATIO:V OF ONE·l\·ODE FREQUENCY 369 Expanding the determinant we get a polynomial equation of nth degree ill (1):! i. e. 0)2 [(I)2(n-1) - Al (I)2(n-2) + A2 (1)~(n-3) - ••• where AI' A 2, A3, ••• , A n - 1 depend on the physical constants of the system, namely the masses and springs. (1)2 = 0 represents a rigid body rotation (not a vibratory motion) and hence can be omitted. There are (n - 1) real values of (1)2. To evaluate for AI' A 2, A 3 , ••• , A n- 1 CROSSLEY and GERl\IEN have devel-oped the following numerical method [1]. Considering a determinant of the form (a1 + b1 - (1)2) - b1 - a2 (a 2 + b2 - (1)2) .J= 0 (u3 - U3 etc. where and 0 0 - b'!. 0 b3 - (1)2) - b3 b =~ n l' n 01 0 0 =0 (12) the solution for the above determinant can be illustrated by the following numerical example of an eight cylinder engine where 11,12, ••• ,18 = 150 lb in. sec'!., 19 = 22.500 Ib in. sec'!., k1' k2' ... ,k7 = 20 X 107 lb in. sec'!., kg = 18 X 107 Ib in.jrad The results are given in Table 1. Col 2. Add the numbers in column 1 from the bottom up, putting the successive two rows higher. Col 3. Multiply the numbers in Column 2 by those in column 1, row by row. Col 4. Add the numbers of Column 3 from the bottom up and set two rows higher, such as for column 2. Col 5. and all odd number columns: :Multiply the numbers of the preceding column by those in Column 1 in the same row. Col 6. and. all even number columns: as with Columns 2 and 4. Lastly, add the total numbers in each odd-number column. These will be the coefficients of the polynomial equation and the equation is 0)2 [W16 - 19.8747 X 106 wH + 159.5885 X 1012 (1)12 - 663.3197 X 1018 W 10 + -+- 1516.3387 X 102.j (i)8 - 18(;6.5223 X 1030 0;6 + 1117.4365 X 1036 (1)4 -- 252.4550 X 1042 (i)2 9.4656 X 1048 ] = 0 (13 370 G. I'. RAM.·n'AIAH Table 1 CROSSLEY'S and GERME:\"S CO!. ljlO' Co1.2jlOfi Col. 3/10" Col. 4/10" Col. 5/101~ Col. 6/10" i Col. 7/10-:: t i Col. S/10" I aI = ko = 0 11 bI kI 17.2076 22.9429 115.4804 153.9701 381.2022 508.2569 639.3792 = - = 1.3333 11 a. k1 1.3333 15.8743 21.1652 96.1129 128.1474 276.5339 368.7026 382.0863 = 1~ b~ = k~ = 1.3333 14.5410 I 1~ ! 19.3675 78.5031 104.6683 192.9745 257.2929 211.2186 k. 13.2077 17.6098 a3 1: = 1.3333 62.6710 83.5593 128.1540 i 170.8677 104.9S31 k b3 = /- = 1.3333 11.8744 15.8321 48.6166 6·1.8206 79.7019 . 106.2655 44.6236 3 k a. = 1: = 1.3333 10.5411 14.0544 36.3399 48.4521 4·5.2483 60.3295 14.7271 b.! k, 1.3333 9.2078 12.2767 25.8408 34,4536 22.4229 29.8965 2.9199 = 1. k a:; = -2 = 1.3333 7.8745 10.4991 15 17.1194 22.825-1 8.8556 11.8072 0.0189 b; k. 6.5412 = / = 1.3333 8.7214 10.1757 13.5673 2.1758 2.9010 , k. 5.2079 6.9437 ar, = J = 1.3333 5.0099 6.6798 0.0142 0.0189 r, br, k., 1.3333 3.8746 5.1660 1.6212 2.161() = ]~ k a, = --'!. = 1. 3333 1, 2.5413 3.3885 0.0106 0.01-12 b, k. = i = 1.3333 1.2080 1.6106 k. as = 1: = 1.3333 0.0080 0.0106 k bs = -~ = 1.2000 1, ag kg ]g = 0.0080 bg kg = 0.0000 = 19 19.8747 159.5885 663.3197 ! 1516.3387 : EVALUATIOS OF OSE·SODE FREQCESCY 371 method to evaluate a large determinant Col. 1i10' CoL9/10'" i CoL 10/10" i CoL 11/10" ' CoL 12/10" I CoL 13/10" I Col. 14/10" Col. 15/10" I 852.48431504.6023 509.4357 1222.9485 i I 281.6178 l 83.0505 i I 139.934·0 ! 23.5539 59.4966 r 3.9183 I k3 I' a., = J; 1.3333 19.6356, 0.0252 k4 b, = -,- = 1.3333 .. , k a;; / = 1.3333 b,; k;; 1.3333 J; k. -Jo = 1.3333 G 1.3333 1.3333 k. [ J~ = 1.3333 i k. J~ = 1.3333 k, 1.2000 J s k ' (/" = J: = 0.0080 k b9 = 19 = 0.0000 3.8931 0.0252 1866.5223 ~~~ -."~-.- --~-~-i ! i I I , I 672.7862 147.3931 1196.5192\ 7.0546 9.4060 36.6619 \ I 297.2572 48.8813 0.0448 0,0596 I I 110,7312 - 2-~- I 7.0098 ;:" ;:'i;:, I I I ! 31.4044 0.0336 0.0448 5.2242 0.0336 9.4656 372 3. Evaluation of the polynomial equation for its roots Equation (13) has eight real roots for co~ leaving co2 = 0 which represents the rigid hody motion as already mentioned. This equation can he solved hy (i) using a computer, or (ii) Graeffe's root squaring method, or (iii) Newton's approximations. Using a computer a series of calculations are to he carried out hy in-creasing co2 automatically hy equal amounts until there is a change of sign. This increment would repeatedly he halved until the required value of co2 is oh-tained with sufficient accuracy. NEWTON'S method of solving the polynomial equation would he illus-trated helow for Equation (13). Let f(x) = x8 - 19.8747 X 106 x7 + 159.5885 X 1012 x6 -- 663.3197 X 1018 x5 + 1516.3387 X 1024 Xi - 1866.5223 X 1030 x1 + + 1117.4365 X 1036 x2 - 252.4550 X 104~ x + 9.4656 X 1048 = 0 where co2 = x Then for a first approximation the lowest root would he X I 9.4656 X 1048 252.455 X 1042 = 3.7494 X 104 !(xl ) :C:: 1.4756 X 1048 If f'(x) is the derivative of Equation (13) with respect to x then x 2 the second nearest approximation for the root is equal to 1.4756 X 1048 :. x2 -'"'-3.7494Xl04 + =4.5868xl04 1.762199 X 1044 where !'(x1) -'"'- - 176.2199 X 1042 This process should he continued until two successive values of the root are the same. GRAEFFE'S root squaring method is dealt with in references [2] and [3]. The process of forming the algehraic equation is, in itself, so time con-suming that these methods cannot he considered quick enough even though they are accurate. EVALUATION OF ONE-NODE FREQUESCY 373 Among those which do not require the frequency equation some are accurate and some approximate. HOLZER'S, matrix, continental, impedance, admittance, mobility and analogous methods can be termed as accurate methods whereas BRADBURY'S, B.I.C.E.R.A.'s, LEWIS', graphical, semigraphical and GUPTA'S methods as approximate ones. Out of the accurate methods HOLZER'S is the simplest and most direct one. But its disadvantage is that the frequency is to be first assumed and then verified whether it is true or not; in other words, it can be said that it is a method of successive approximations. So if the first assessment can be made quickly and accurately then HOLZER'S calculations become very easy and hence the approximate methods should be resorted to in finding the nearest one-node natural frequency. 4. Lewis' method [4] When the approximate position of the node of the system under consid-eration is kno·wn, then this method can be used and is less time consuming. Accordingly (J) = --- ~ rad/sec ;-z; [k ]112 2 J E where J E = E Jey !> Jey! = moment of inertia per line, key! = crankthrow stiffness (between cylinder centers) and the node is situated close to the fly-wheel. co = -~-J 20 X 107 1112 2 L 8.5 X 1200 = 35 cy/sec rad/sec, (15) This formula always gives a higher value compared to most of the other methods because of the assumption that one node is at the fly-wheel. If a more accurate value could be assessed by using any other approximate method and without needing more time then it would be more practicable. 5. B.I.C.E.R.A. formula [4] According to this [ k Jl/2 [ 1 I B ]1/2 F=9.55 ~ .: T J ey! AB cy/min (18) 374 G. re RAMAiXAlAH where keyl : crankthrow stiffness Ib in./rad, J eyl : moment of inertia per cylinder Ib in. sec~, B: ratio of flywheel inertia to moment of inertia per line, A : N(N 1)/2 where N is the number of cylinders. F 9.55 X 103 (200)1/2 (186)112 = 34.11 cy/sec 900x 60 .. Assumptions made in the derivation of the formula are that the node position is- at the flywheel and the relative amplitudes" decrease linearly from the cylinder end to the flywheel end. Both assumptions are not true (even though assumed to be true for a first approximation) and hence the discrepancy in the value obtained for the one-node natural frequency. 6. Bradbury's diagram for estimating the one-node natural frequencies [5] BRADBURY'S diagram to find one node natural frequencies is plotted between the ratio (keydJe) I X 106) on X-axis (frequency in vib/sec), on Y-axis (number of cylinders as parameter). keyl represents the stiffness per crankthrow and Jcyl the moment of inertia per line. The assumptions made while plotting the graph are (i) all the moments of inertia per line are equal, (ii) the ratio J pjJeyl is not less than 40 (J F = moment of inertia of the flywheel), (iii) all crankthrow stiffnesses are equal and the stiffness of the shaft portion kF between the end cylinder cent er and the fly-wheel is also equal to keyl ' The frequency value could also be calculated from the formula F = uv(kc .. )1/2 ----:.)-, cv/sec. 103 (JCy\)1/2 ., .. (17) where u = 56.9,46.9,40,34.9,31.05,27.9 and 25.45 for 4, 5, 6, 7, 8, 9 and 10 cylinders respectively, and v the correction factor, has been given by BRADBURY in the form of another graph; number of cylinders on X-axis, v on Y-axis with keydkF as parameter ranging from 0.8 to 1.2. Results obtained are not true to a reasonable extent (as would be shown. later) because of the inaccuracy in the second assumption above. F = 31.05 X 0.989 X 10:1 [4/3]1;2 103 35.459 cv/sec • I EVALUATIOS OF OSE-SODE FREQl-ESCY 375 7. Gupta's method [6] Charts h n-e been prepared by GUPTA by matrix evaluation to find the fundamental natural frequency and relative amplitudes. Understanding the procedure requires defining some terms as and with the usual notation. For the problem for which the solution is required x 0.0066, fJ = 1.1111. Referring to Fig. 6, page 163 [Ref. 6], x2 = 0.0345, F = _1_ (keYJ x,x )1/2 . cyjsec, 2;;: I eYJ . = 34.139 cy/sec, (18) (19) (20) and the relative amplitudes can be calculated from Fig. 9, page 164 [Ref. 6]. According to BRADBURY, OJ is the same for all engine systems if the value of ca is less than 1/40. But following GUPTA'S charts, x 2 = 0.0335 for x = 0.001, = 0.0388 for x = 0.025, and F = 33.637 for x = 0.001, = 36.203 for :x = 0.025. So BRADBURY'S diagram does not give correct values of ca even for values of IF/leyJ more than 40. BRADBURY [7] is also the first to give data facilitating the construction of GORFINKEL tables, for engines having five to eight cylinders, with equal moment of inertia per cylinder line and stiffness equal between cylinders. The variation of flywheel inertia also is taken into consideration. The values of relative amplitudes and x (as indicated in GUPTA'S charts) can be Tead off for any orthodox system and the natural frequency can be calculated by inserting stiffness and moment of inertia values. GORFINKEL'S method [8] is in itself a simplified modification of HOLZER'S, 'which reduces the arithmetical -work for engine systems comprising a large number of cylinder inertias JeyJ and crankthrow stiffness keyL with identical values. 4 Periodica Polytechnica ~I. 12j4. 376 G. F. RAMAXAIAH It is not out of place to mention that in Engineering of 19th Feh~uary, 1937 there is an article that gives a formula for the lowest natural frequency of any system with four to ten cylinders, values of Cl; ranging from 0 to 0.04 and fJ from 0.8 to 1.2. 8. Other methods Matrix methods do not offer any advantage over HOLZER'S method or solution of the determinant equation as far as simple systems are concerned, except for the fact that they are more useful for solution of complex systems and can he readily programmed for a digital computer. Graphical and semi-graphical methods takc comparatively more time and the accuracy depends mainly on the accuracy of the diagram. Impedance, admittance, mobility and analogous methods are easy for simple systems but hecome complicated when multimass systems are dealt with. With methods described so far the assumptions are that the equivalent system constitutes weightless springs and point polar mass moments of iner-tia. In the distributed mass method and, to some extent, in the continental method for frequency evaluation, the masses of springs are also considered. But they involve more time in calculating the one-node natural frequency. From all these considerations it could be concluded that the natural frequency of the engine system under study could be obtained 'within the least amolmt of time and with maximum possible accuracy by using GUPTA'S charts and HOLZER'S method combined. 9. Comparison Comparison of the one-node natural frequencies obtained hy different methods shows that GUPTA'S charts give the nearest approximate value to the correct one. 31ethod BRADDt::RY Frequency 35.459 10. Conclusions 1. Different methods to evaluate the one-node natural frequencies are compared to arrive at the best possible way of determining the one-node EVALUATIOS OF OXE.SODE FREQUe:\"CY 377 natural frequencies of multicylinder engine systems with improved rapidity and accuracy. 2. It has been found that the best approximate method for one-node natural frequencies and angular amplitudes is given by GUPTA. 3. A more accurate assessment can only be made by HOLzER's calcula-tions. n. Summary ~umerou5 methods for evaluating the one-node natural frequencies of multicylinder engine systems are available in technical literature. As such an attempt, this paper suggested the course to be taken in determining the one·node natural frequencies and angular amplitudes with improved rapidity and accuracy. 12. Acknowledgements The author thanks Prof. Dr. A. Jurek for the encouragement given during the preparation of the paper, and other colleagues of the department of Gas Engines and Automobiles of Buda-pest Poly technical University - for the help rendered. 13. References 1. CROSSLEY, GER:'lIE:'>: Method of numerical evaluation of a large determinant, Jonrnal of Applied Mechanics, June 1960. 2. VON K..iR:>IAN, TH., BIOT, M. A.: Mathematical methods in engineering, Mc Graw-HilI Book Co .. New York 1940. 3. HANSE:'>, H.'M., CHE:'>EA, P. F.: Mechanics of vibration, John Wiley and Sons, Inc., New York 1952. 4. XESTORIDES, E. J.: B.I.C.E.R.A. Handbook on torsional vibration, Cambridge University Press, 1958. 5. BRADBURY, C. H.: Torsional vibration in Diesel engines, Charles Griffin and Co., Ltd. London, 1938. 6. GUPTA, K. N.: Matrix techniques for determination of fundamental mode shape and fre-quency parameter of torsional oscillations in engines, J otlrnal of Mechanical Engineering Science, 4, 1962, 156-165. ,. BRADBURY, C. H.: Torsional vibrations in Diesel engines: Some observations and practical aspects, Diesel Engine Users' Association, Paper No. S. 226, l\Iarch 1953. 8. GORFINKEL, A.: Critical speeds of crankshafts', Engineering, 27 December 1929, pp. 827-29. 9. DEN HARTOG. J. P.: Mechanical vibrations, 2\lc Graw-Hill Book Company, Inc., 1956. 10. WILSO:'>, W. K.: Practical solution of torsional vibration problems, Vols. 1 and 2. Chapman and Hall Ltd. London, 1963. 11. WILSON, W. K.: Vibratio~ engineering, Charles Griffin and Co., Ltd., London, 1958. 12. BISHOP, R. E. D., JOHNSON, D. C.: The Mechanics of Vibration, Cambridge University Press, 1960. 13. THO)ISON, W. T.: 2\Iechanical vibrations, G. Allen and Unwin, Ltd., London 1950. G. V. RAl'tiAl'AIAH, Budapest XI., Sztoczek u. 2, Hungary. 4* 

ON THE ACCURACY AND RAPIDITY OF EVALUATION OF ONE·NODE NATURAL FREQUENCY AND ANGULAR AMPLITUDES OF UNDAMPED MULTICYLINDER ENGINE SYSTEMS

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ON THE ACCURACY AND RAPIDITY OF EVALUATION OF ONE·NODE NATURAL FREQUENCY AND ANGULAR AMPLITUDES OF UNDAMPED MULTICYLINDER ENGINE SYSTEMS

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