A minimized mechanical impedance is desirable in order that a phonograph pickup can trace a sound groove correctly. Current phonograph pickups, however, have resonance peaks in their frequency responses and mechanical impedance characteristics in the neighborhoods of their high frequency ends. This has introduced a great problem to sound groove tracing. Current pickups have high resonance frequencies beyond the hearing limit rather than damped resonances which can be achieved by lowering the equivalent mass of the vibrating system as shown in Fig. 2. But this method gives some problems such as degradations of sensitivities and fragilities. In this paper, as an effective method for damping out the high frequency resonance, a vibrating system with two degrees of freedom has been studied. As a result of the analysis, it has been made clear that the following method is useful. That is splitting the cantilever into two parts and connecting them with elastic material with small resistance, and damping the high frequency resonance by the resistance of the armature bearing. The split cantilever makes not only the high frequency flat but also the mechanical impedance characteristic in the high frequency range flat, as shown in Fig. 6, therefore the split cantilever is suitable for a high quality phonograph pickup. Experimental phonograph cartridges with split cantilevers (Fig. 7) have been made. Fig. 8, shows the frequency response and the mechanical impedance characteristic of the experimental cartridge. It's results agree well with the calculated results shown in Fig. 9.
This paper deals with analyses and some experimental results on the free-free "steeped bar" which is stepped at one side or both sides of a uniform bar(Fig. 1). It is the purpose of this paper to obtain a guide for the design of a small sized resonator in low frequency. The stepped bar can be produced easily and can by supported at the nodal points which appear in the thick parts of the bar. These results are also useful for the design of a tuning fork. First, the fundamental mode vibration of the bar stepped at one side (Fig. 2) is analyzed with the mechanical equivalent network (Fig. 3), and the frequency-equation is obtained (Eq. 9), and the equation of the vibration mode is given (Eqs. 16, 21). The calculated and measured values of resonant frequencies are shown (Figs. 4, 5), and the resonant frequencies of the bar stepped to the thickness direction are compared with the ones of the bar stepped to the width direction (Fig. 6). The frequencies of higher modes are also shown, in (Fig. 7). It is clear from the above results that as the cross section ratio of the bar increases, the more the resonant frequency decreases, and that the minimum resonant frequency is obtained at the length ratio of 0. 1〜0. 35. Second, the fundamental vibration mode and its nodal point of the bar stepped at one side are calculated and compared with the measured values (Fig. 8). It is confirmed that the bar can be supported at a nodal point of vibration in the thick part of the bar. Lastly, an analysis on the bar stepped symmetrically at both sides is summarized and the equations of resonant frequency and vibration mode are obtained (Eqs. 25, 28, 31), and the design charts of the bar are given (Figs. 10, 11, 12). The resonant frequencies of the bar stepped at both sides come down to about 0. 7〜0. 85 times the ones of the bar stepped at one side. Accordingly, the two stepped bar is more suitable for a small size resonator in low frequency than the one stepped bar.
This paper deals with the silencing effect of exhaust mufflers of automobiles. The purpose of this experiment is especially to investigate the effect of the flow rate of exhaust gas on a changes in noise attenuation in mufflers from the practical point of view. Ten fundamental types of reactive mufflers are tested by an engine, which has a cylinder volume of about 900 cc. The results are compared with those of a test by a pulsating air flow at various flow rates. The spectrum of noises in the pulsating flow test is similar to that of exhaust noises by the engine test as shown in Fig. 1. But in general, the sound level in the pulsating flow test is lower in the frequency range, but higher in the high frequency range than the level in the engine test (Fig. 2). The noise attenuation L_e (Eq. 1) in the engine test varies with engine load, and is generally in value under the full load than in no load condition, especially in the high frequency range above about 300 Hz (Fig. 4). These tendencies are observed clearly in the case of the mufflers with the inlet and outlet pipes installed straight through. The noise attenuation L_p in the pulsating flow test also decreases with an increase in flow rate (Fig. 4). With the noise spectrum divide into two frequency ranges, the low and high frequencies bounded by 500 or 1, 000 Hz, the acoustic power of the exhaust noise W' is classified in thems of a function of the mean flow rate as shown in Eq. 2〜5. The acoustic powers, W_L and W_H, of exhaust noises from an exhaust pipe without a muffler consist of the components proportional to the 2nd, 4th or 6th power of the mean flow rate as shown in Fig. 5. And the acoustic power W_m of noises generated in a muffler can be explained in terms of the component proportional to the 6th power of the flow rate (Fig. 6). The noise attenuation in a muffler is expressed as shown in Eq. 8. The amount of attenuation L in a muffler is comparable between the engine test and the pulsating flow test under the condition expressed by Eq. 10 (Fig. 9). In this case the coefficient of correlation between the results of both tests is given as about 0. 78 for the whole mufflers (Fig. 10).