vibracion diesel engines.pdf

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Engine dynamics and vibration control AU T H O R S : H a n n u T i e n h a a ra , H e a d o f C a l c u l a t i o n & S i m u l a t i o n , R & D, W r t s i l i n F i n l a n d H e i k k i M i ko n a h o, S t r e n g t h & S t r u c t u r a l D y n a m i c s , R & D, W r t s i l i n F i n l a n d

Fig. 1 A Wrtsil 8L46 engine with ABB TPL turbocharger.

The increasing demand for lowering the noise and vibration levels of engines has forced manufacturers to make use of advanced analysis and simulation tools.

In most cases, the practical means to reduce vibration is simply to detune the lowest natural frequencies away from the main dynamic excitation frequencies. When detuning natural frequencies, the most effective course is to concentrate on the heavy structures built on to the engine and its mounting. A good example is the turbocharger, because its inuence on the vibration system is very dominating due to its relatively large mass (Figure 1).

In certain problematic situations, a tuned mass damper can be used to change the vibration system dramatically.

As regards reducing vibration on the Wrtsil 9L46 engine, a study ended

up with two different solutions: For the current production engines, a new ring order was introduced offering a better distribution of the excitation forces at certain harmonic orders. This solution requires the use of a special balancing device in order to cope with the increased rst order free couples. However, changing the ring order on a 9-cylinder engine is not a feasible solution for existing engines already in the eld. For these engines the tuned mass damper was chosen as being the most suitable solution.

A tuned mass damper is a device whereby an additional mass is mounted with exible elements on the vibrating machine. The damper is tuned in such a way that its own vibration is producing a counter force against the main structures vibration. Normally a damper is tuned to dampen a certain natural frequency, but in the case of a

constant speed engine, it can also be tuned to a specic excitation frequency.

The two biggest challenges in designing and tuning this kind of a system are: 1)Handling a wide range of running

speeds and several natural frequencies and mode shapes.

2)Making a reliable construction capable of operating for thousands of running hours without maintenance.

Dynamic system The relationship between the excitation, the structural properties, and the response can be expressed as per the diagram in Figure 2. The vibration response is a result of the dynamic properties of the structure and the excitation force.

A vibration system is normally presented mathematically by the well-known general equation of motion:

(1)

where M, C and K are matrices of mass, damping and stiffness, f(t) is the vector of applied force (excitation), and x(t) is the vector of displacement (response) and its time derivatives, velocity and acceleration accordingly. The matrices M, C and K represent the dynamic characteristics of the structure. Reducing vibration levels can be achieved by modifying one or several of these characteristics, or the excitation vector f(t).

The matrix M is not only the total mass, but represents also the mass distribution over the whole structure. The same applies to the stiffness matrix K. From the vibration point of view, it may be very important where the mass or stiffness is located.

C denotes the damping, which in practice is not only a uniform number. In real structures the damping normally varies depending on the frequency and mode shape, as well as on the location. In complex structures like engines, several different damping types can be found.

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Finally, there is the vector f(t) representing the force or excitation vector. Means of modifying the force vector in order to reduce vibration response can be, for example, a balancing device where some additional forces are included in the system, or changing the ring order when the forces are applied in a different order to the system.

Vibration analysis In the vibration analysis of an engine or a diesel generating set (genset), the following parts can be included: 1) Eigenfrequency and mode shape analysis 2) Calculation and analysis of major excitation forces 3) Dynamic response analysis.

Eigenfrequency analysis Normally, the rst step in making a structural vibration analysis of a diesel engine is a calculation of its natural frequencies and mode shapes.

A typical calculated lowest torsion mode is shown in Figure 3.

Excitations The calculation and analysis of excitation forces are an essential element in vibration optimization. The major excitations caused by gas and mass forces, taking into account the ring sequence of the engine, are calculated and analyzed. Modern multi-body dynamics (MBD) simulation tools offer an accurate and relatively fast way of calculating the mechanical excitation forces acting on the engine block.

The excitations of diesel engines are periodic. For this reason, it is natural to analyze the excitations as well as the vibration measurements within the frequency domain.

The main excitation sources of a medium speed diesel engine can be categorized as shown in Figure 4, [1]. The origin of mass forces is the crank mechanism, which has both rotating and oscillating components.

On the lowest integer harmonic orders, the mass forces induce mainly rigid body motions of the whole engine structure. However, some bending of the engine block due to mass forces, is also visible, especially on long engines. The gas forces resulting from the cylinder pressure cause a torque variation at each cylinder. This torque variation is transferred to the engine block through the main bearings and via the lateral force

Fig. 2 Diagram of the relationship between excitation, structure and response.

Fig. 3 Typical lowest natural mode shape of torsion of a Wrtsil 8L46 engine.

Fig. 4 The main categories of excitation forces.

Excitation Frequency, amplitude, direction, location, etc.

Structural properties Natural frequencies, natural mode shapes, damping

Vibration response Amplitude, frequency, mode

Excitation type

Excitation source Oscillating

Appearance in a multi-cylinder engine block

Vibration at the rst harmonic order

Rigid body vibration and bending

Vibration at lowest full orders, mainly orders 1 and 2 Rigid body, bending and some torsion

- All harmonic orders, including half orders

- Mainly torsion based de ections

on the engine

SIMPLE ..................................................... COMPLEX

Main excitations of a 4-stroke engine

Gas forces

Cylinder pressure Rotating (Unbalance)

Mass forces


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Fig. 5 Relative comparison of torsional excitations of a 9-cylinder engine with two ring orders.

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Fig. 6 Inuence of an added spring-mass-damper system on the resonance frequencies and response amplitudes of the main structure.

Fig. 7 The prototype of the tuned mass damper.

of the piston against the cylinder liner. When analyzing these excitations, it

is necessary to take into account, not only the excitation strength at different frequencies, but also the similarity of the excitation mode and natural mode shapes within the frequency range in question.

As regards gas-force-induced torsion excitations, it is relatively easy to quickly compare different ring orders by means of a vector summation. Figure 5 shows the difference in torsion excitations with two different ring orders on a 9-cylinder engine.

Forced response analysis When using MBD software to simulate engine vibrations, the simulation model itself performs the calculation of the excitation forces and their application on the correct locations in the model. The analysis is done in a time domain, normally using condensed models of the structure.

The direct time integration method, using Finite Element Software, is very time consuming and often not feasible. Its most important advantage is that it can take the structural nonlinearities into account. A linear analysis in a frequency domain is fast and sufciently accurate providing that the FEM model is presenting the structural characteristics reliably.

Tuned mass damper The use of a passive tuned mass damper is a known method for reducing vibrations resulting from earthquakes in high buildings. It has also been used to eliminate vibration problems on ship structures, for example, and to solve different kinds of machinery vibration problems, but not necessarily so much for reducing diesel engine vibrations.

The designation tuned mass damper refers to the construction, consisting of a vibrating mass with a natural frequency, tuned to the desired frequency. Figure 6 shows the principal effect of a tuned mass damper. The device is also known as a vibration absorber in situations where the damping factor is very small, such as when just a steel spring is used without any additional damping.

The blue line shows the vibration response of the main structure m0 due to ground movement or applied force. It has a natural frequency at frequency f0 where the increased vibration amplitude can

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be seen. The red line shows the response of the same main structure after adding mass m1 to the system with the spring k1 and damping c1, as shown in the Figure. When the damper is correctly tuned, it can reduce vibrations dramatically within the area of the resonance frequency.

From Figure 6 it can also be seen that by adding a damper to the system, the vibration of the main structure outside the intended damping area increases. This is one disadvantage of the damper, which must be taken into account. One must remember, particularly in the case of medium speed diesel engines where the main excitation frequencies are spread over a wide frequency range, that at some harmonic orders the vibration level is increased by the damper. However, with proper tuning of the mass, stiffness and damping parameters, it is possible to reduce this phenomenon.

As it has been clearly shown, the correct dimensioning of the added mass m1, the stiffness k1 as well as the damping c1, is essential in order to achieve the best possible damper performance. As a rule of thumb, it can be considered that the added mass required to achieve a proper damping effect, is about 5% of the modal mass of the vibration mode in question. The spring coefcient k1 and the damping factor c1 are then chosen so that the damped natural frequency of the mass m1 will match the frequency to be dampened.

Engine vibration control by using a tuned mass damper Contrary to the above theoretical example,

in the case of a real engine, the problem is somewhat more complicated. Firstly, nding the correct parameters is not an easy task when the engine has a wide range of rotating speeds and several harmonic orders exciting resonances. Secondly, nding theoretically the best possible location and direction for the damper requires a thorough analysis of the system using the nite element method. Thirdly, actual structures usually consist of several natural mode shapes that contribute to excessive vibration levels. By choosing a suitable direction for the damper, it is still possible to have some inuence on more than one mode shape.

Damper development The tuned mass damper developed by Wrtsil consists of vibrating mass discs supported by steel springs. Both are located, together with damping oil, inside a cylindrical steel frame. All the damper parameters, mass, stiffness and damping, can be separately adjusted. The damping coefcient is changed by altering the oil ow inside the damper. The damper is shown in Figure 7.

Vibration simulation and tuning Comprehensive simulations were carried out during the development of the mass damper. The main parts of the engine model were built in Ideas, and the meshing was done in Hypermesh. The engine vibrations with the tuned mass damper were simulated using Abaqus and Modysol software.

Modysol is a software package developed

by VTT, the Technical Research Center of Finland.

The dynamic excitation forces were calculated using an in-house software called Dynex.

Optimizing the location of the damper is essential to minimizing its effective mass. After several simulations it was noticed that, in the case of an in-line Wrtsil 46 engine, the top of the turbocharger is the most feasible location in order to minimize the required vibrating mass.

A 9-cylinder four-stroke engine with a ring order of 1-2-4-6-8-9-7-5-3, gives high excitation forces for the rst torsion mode at harmonic orders 4.0 and 5.0, as shown in Figure 8. Between those excitations, the harmonic order 4.5, which corresponds to the ring frequency, gives a strong rolling excitation.

With the damper mounted so that the movement of the effective mass is in the engine transversal direction, it is not very efcient in damping the vibrations in the vertical direction. According to vibration measurements, the vertical vibration is, however, not very critical in this case. It is clear, therefore, that there are two possibilities for tuning the damper: to concentrate on the 1st torsion mode at 29 Hz, or on the horizontal bending mode at 42 Hz. The former option was proven to be the better one, and was nally chosen.

The same principles are used in the tuning of the damper for the Wrtsil 8L46 engine. The most signicant excitations are at the harmonic orders 3.5 and 4.5, as well as at the order 4.0, which corresponds to the ring frequency of an 8-cylinder engine.

Fig. 9 A tuned mass damper mounted on a turbocharger.

Fig. 8 The most critical natural frequencies and excitations for the 1st torsion mode on a Wrtsil 9L46 engine with a ring order of 1-2-4-6-8-9-7-5-3.

2.0 2.5 3.0 3.5 4.0 4.5 5.0 5.5 6.0

Natural frequencies


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Fig. 10 Reduction of vibration levels on the turbocharger using the tuned mass damper. Wrtsil 9L46 engine.

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Fig. 11 Comparison of overall vibration levels on a Wrtsil 9L46 engines turbocharger and silencer with and without a mass damper.

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Field testing and results In addition to comprehensive simulations on both the 8-cylinder and 9-cylinder Wrtsil 46 engines, the mass damper has passed full scale eld testing on real engines, as shown in Figure 9.

The eld testing was carried out in order to verify the functionality of the damper with the optimum tuning parameters, as well as to assess the performance over long term operation. The reduction of the vibration levels on the turbocharger compressor casing are shown in Figure 10. From these measurements it can be seen that with the damper well tuned, the vibration can be reduced at more than one harmonic order. In this case, vibrations at all the three major excitation harmonics are reduced in both transversal and longitudinal directions. Vibration is increased only at the harmonic order 4.0 in the vertical direction, but also there the overall vibration level is slightly reduced. This can be seen in Figure 11 where L, T and V denote the longitudinal, transversal and vertical directions, respectively.

Figure 12 indicates the overall r.m.s. velocity vibration levels on the same turbocharger during the engine sweep run. The running speed was changed from 360 up to 500 rpm with a propeller loading.

Another example, shown in Figures 13 and 14, is taken from the vibration test results of an 8-cylinder engine. The gures show the results, with and without the tuned mass damper, on the engines foot, charge air cooler, and turbocharger (Figure 14).

At the writing of this article, altogether 15 dampers have been delivered. The cumulative running hours for these dampers is 82,000 h. However, some of these dampers have already accumulated 12,000 running hours without any service operation.

CONCLUSION When optimizing the vibration performance of a medium speed diesel engine, as many different contributing aspects as possible should be taken into account. A thorough vibration analysis includes the eigenfrequency and mode shape analysis, the analysis of excitation forces, and nally, as a combination, the dynamic forced response simulation

The nal result is always a compromise between many different criteria. For example, the ring order giving the

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smallest free forces is most probably not the best one from the point of view of the internal bending moment or torsional vibration. Similarly, the stiffening of the structure in order to move one natural frequency away from a critical excitation, may create another natural frequency in another excitation frequency area.

When it is not possible to tune the natural frequencies of the engine structure properly to avoid vibrations, and when modifying the excitation forces is not feasible, a tuned mass damper can be a good solution.

An accurate prediction of the performance of a tuned mass damper requires special simulation tools, making it possible to include the local damping within the simulation model.

On the basis of the simulations and tests described in this article, the best and most effective location for the tuned mass damper is on the turbocharger. At that location the displacement amplitudes are normally much higher than on other parts of the engine, which is essential for the damper to work ef ciently.

When the damper is properly tuned, it can reduce vibration levels at more than one excitation frequency. In the cases presented here, considerable vibration reduction was achieved at three major excitation harmonics.

NOMENCLATURE C Damping matrix M Mass matrix K Stiffness matrix f(t) Force vector x(t) Displacement vector L Longitudinal T Transversal V Vertical c0 Damping of the main structure

mounting c1 Damping of the damper mass mounting k0 Stiffness of the main structure

mounting k1 Stiffness of the damper mass mounting m0 Mass of the main structure m1 Added damper mass r.m.s. Root Mean Square

REFERENCES [1] TIENHAARA, HANNU, Guidelines for engine dynamics and vibration, Wrtsil Marine News, 1/2004

Fig. 12 Vibration on the turbocharger in the vertical direction. Wrtsil 9L46 engine.

Fig. 14 Vibration results on an 8-cylinder engines turbocharger with and without a damper.

Fig. 13 Vibration results on an 8-cylinder engine with and without a damper.

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vibracion diesel engines.pdf