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International Journal of Turbomachinery, Propulsion and Power
, Volume 7 (2) – May 6, 2022

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International Journal of Turbomachinery Propulsion and Power Article Acoustoelastic Modes in Rotor-Cavity Systems: An Overview on Frequency Shift Effects Supported with Measurements Tina Unglaube * and Dieter Brillert Chair of Turbomachinery, University of Duisburg-Essen, 47057 Duisburg, Germany; dieter.brillert@uni-due.de * Correspondence: tina.unglaube@uni-due.de Abstract: With an increase in ﬂuid densities in centrifugal compressors, ﬂuid-structure interaction and coupled acoustoelastic modes receive growing attention to avoid machine failure. Besides the vibrational behavior of the impeller, acoustic modes building up in the side cavities need to be understood to ensure safe and reliable operation. In a coupled system, these structure and acoustic dominant modes inﬂuence each other. Therefore, a comprehensive overview of frequency shift effects in rotor-cavity systems is established based on ﬁndings in the literature. Additionally, experimental results on coupled mode pairs in a rotor-cavity test rig with a rotating disk under varying operating conditions are presented. Measurement results for structure dominant modes agree well with theoretical predictions. The development of a forward and a backward traveling wave is demonstrated for each mode in case of disk rotation. Conducted experiments reveal the occurrence of weakly and strongly coupled mode pairs as frequency shifts are observed that cannot solely be explained by “uncoupled mode effects”, such as the added mass, speed of sound, and stiffening effect, but indicate an additional coupling effect. However, the hypothesis of a bigger frequency shift for stronger coupled modes cannot be corroborated consistently. Only for the strongly coupled four nodal diameter mode pair in the “wide cavity” setup, a coupling effect is clearly visible in the form of mode veering. Citation: Unglaube, T.; Brillert, D. Keywords: rotor-cavity system; acoustoelastic modes; coupling; ﬂuid-structure interaction Acoustoelastic Modes in Rotor-Cavity Systems: An Overview on Frequency Shift Effects Supported with Measurements. Int. J. Turbomach. Propuls. Power 2022, 7, 15. https:// 1. Introduction doi.org/10.3390/ijtpp7020015 Centrifugal compressors are applied for a wide range of industrial applications, in- Academic Editor: Antoine Dazin cluding gas compression for the chemical industry, natural gas extraction and liquefaction, enhanced oil recovery, as well as carbon capture and storage. Required discharge pressures Received: 26 January 2022 increase continuously, and so do ﬂuid densities inside the machines, increasing the forces Accepted: 4 May 2022 acting on the impeller. Besides the vibrational behavior of the impeller, which has been Published: 6 May 2022 studied extensively over the last decades, these ﬂuid-structure interactions have to be Publisher’s Note: MDPI stays neutral understood. Nowadays, it is well known that natural frequencies of compressor impellers with regard to jurisdictional claims in shift due to an increase in gas density of the surrounding ﬂuid. Moreover, acoustic modes published maps and institutional afﬁl- building up in the side cavities of radial compressors were found to be another potential iations. source of high cycle fatigue. Recent studies suppose that acoustic and disk vibration modes inﬂuence each other, resulting in further frequency shifts in coupled rotor-cavity-systems. However, these effects are not fully understood yet. To ensure a safe and reliable de- sign and avoid machine failure during operation, these coupled natural frequencies have Copyright: © 2022 by the authors. to be predicted accurately, and the excitation and the damping mechanism need to be Licensee MDPI, Basel, Switzerland. studied further. This article is an open access article Rotor-cavity systems, consisting of a rotating disk in a ﬂuid-ﬁlled cavity, provide a sim- distributed under the terms and pliﬁed model of centrifugal compressors for fundamental research purposes. Therefore, a conditions of the Creative Commons comprehensive overview on frequency shift effects in acoustoelastic rotor-cavity systems is Attribution (CC BY-NC-ND) license established based on ﬁndings published in the literature. This provides a robust foundation (https://creativecommons.org/ for a theoretical understanding of the topic, which is then underpinned with measurement licenses/by-nc-nd/4.0/). Int. J. Turbomach. Propuls. Power 2022, 7, 15. https://doi.org/10.3390/ijtpp7020015 https://www.mdpi.com/journal/ijtpp Int. J. Turbomach. Propuls. Power 2022, 7, 15 2 of 21 data. For this purpose, a rotor-cavity test rig was set up at the Chair of Turbomachinery at the University Duisburg-Essen (Germany) that demonstrates the occurrence of coupled acoustoelastic modes. It enables the researchers to systematically collect and evaluate experimental data on rotor-ﬂuid interaction and the formation of coupled acoustic and structure dominant modes while varying different inﬂuencing parameters independently from each other. By doing so, a deeper understanding of coupling and frequency shift effects in rotor-cavity systems is gained. 2. Literature Overview Some of the earliest works on disk vibration in turbomachinery were conducted by Lamb and Southwell (1921) [1] and Southwell (1922) [2], analyzing the vibrations of a spinning circular disk. A few years later, Campbell (1924) [3] published his fundamental study on the vibrational excitation of rotating turbomachinery, thereby establishing the so-called Campbell diagram. Until today, it remains a widely-used tool to avoid potential vi- brational resonances that might result in machine failure. Subsequently, many studies were conducted analyzing the vibrational behavior of rotating disks and compressor impellers. E.g., Eversman and Dodson (1969) [4] published a technical note on the free vibration of spinning, centrally clamped circular disks, and presented an analytical solution of the Eigenvalue equation as a function of hub-to-disk ratio for different spin rigidity parameters and mode shapes. Ewins (1973) [5] published an analytical and experimental study on the vibration characteristics of tuned and detuned bladed disk assemblies, emphasizing the importance of considering the complete bladed disk assembly, instead of disk and blades separately. Irretier [6,7] studied the natural and forced vibrations of a wheel disk and established a numerical model to calculate the Eigen frequencies and mode shapes of the system and later carried out experiments and calculations on the vibrations of rotating radial impellers, which were in good agreement. One of the ﬁrst fundamental studies on aero-acoustic excitation sources in turboma- chinery was published by Tyler and Sofrin (1962) [8]. They identiﬁed rotor-stator interaction as a signiﬁcant noise generating mechanism and found that both the spinning rotor alone and rotor-stator interaction produce rotating pressure patterns. Since then, these spinning modes are referred to as Tyler-Sofrin modes. Ehrich (1969) [9] derived an analytic model of an annular acoustic cavity to conclude on natural frequencies of acoustic modes in the cavity. It is used to analyze experimental data of a turbomachine over a range of rotational speeds and explain resonance peaks at supersonic and subsonic wheel speeds. Amongst the more recent publications, Eisinger (2002) [10] and Eisinger and Sullivan (2002) [11] address acoustically induced fatigue of impellers of rotating machinery. Coupling of the structural and acoustic modes is discussed that might ultimately lead to structural failure of the impeller. This becomes especially important for impellers in high-density ﬂuids, where disk vibration and acoustic modes cannot be treated separately anymore but need to be analyzed as a coupled system. Magara’s publications [12,13] are of particular interest along this chain of causation as they discuss structural frequencies of radial compressors in a high-density environment. Two analytical models are established to explain the sub- sequent frequency shift phenomena for a non-rotating and a rotating disk due to mode coupling. Experimental results agree well with numerical ones. Beirow et al. (2015) [14] studied the effect of mistuning on a high-pressure compressor blisk at rest with respect to structural vibration modes and damping ratios. The forced response is analyzed, and it is found that, taking aero-elastic coupling effects into account, damping of the struc- ture generally increases while the response decreases. Based on other theoretic models, Heinrich et al. (2020) [15] proposed a generalized model for the approximation of coupled acousto-mechanical natural frequencies in high-pressure centrifugal compressors, which was later extended to take damping effects into account (Heinrich et al. (2021) [16]). In the latter, the model outcome is further compared to experimental data, yielding excel- lent agreement regarding the predicted natural frequencies and an acceptable agreement regarding damping ratios. Int. J. Turbomach. Propuls. Power 2022, 7, 15 3 of 21 Examples of aero-acoustic excitation in industrial turbomachinery are, e.g., reported in Hellmich and Seume (2006) [17], where an acoustic resonance in a four-stage axial compressor is explained with a simpliﬁed model for helical acoustic modes, as well as König (2009) [18] and Petry et al. (2012) [19], who investigate acoustic Eigenmodes in the side cavities of an industrial centrifugal compressor. Their setup includes one complete compressor stage, from the inlet guide vanes to the return guide vanes. Another case of actual high cycle fatigue in an industrial compressor possibly caused by aero-elastic excitation is described in Eckert (1999) [20], accentuating that the problem of ﬂuid-structure interaction in turbomachinery is of substantial concern in real applications. The current paper is based on the preceding work of the research group at the Chair of Turbomachinery at UDE, aiming at gathering baseline experimental data on rotor- ﬂuid interaction. Barabas et al. (2015) [21] introduced the test rig and measurement instrumentation in detail and presented ﬁrst velocity proﬁle measurements, indicating a turbulent ﬂow with high Reynolds numbers inside the cavity. Barabas et al. (2017) [22] presented frequency shifts towards higher and lower frequencies for the two modes of a coupled mode pair for a change in cavity pressure. Moreover, interfering inﬂuences of crossing modes were identiﬁed. Barabas et al. (2018) [23] then focused on the damping of weakly coupled acoustic dominant modes, which was found to decrease with an increase in ﬂuid pressure, and established a relationship of the damping ratio, the kinematic viscosity, and the natural frequency of the acoustic modes. It is shown that structural-acoustic mode coupling in rotor-cavity systems is a complex ﬁeld of research with many interacting effects to be taken into account. Several publications address some of these effects, but so far none has revealed them at large. This paper contributes to a deeper understanding of the coaction of all pertinent frequency shift effects in rotor-cavity systems and thus to the efforts to avoid resonances in centrifugal compressors that might ultimately lead to component failure in turbomachinery. 3. Theory on Frequency Shifts in Rotor-Cavity Systems 3.1. The Rotor-Cavity System A rotor-cavity (or rotor-stator) system consists of an impeller rotating inside a station- ary cavity. It can be used as a simpliﬁed model of turbomachine impellers in a casing. The rotating disk divides the ﬂuid volume into two separate cavities, similar to the side cavities in, e.g., centrifugal compressors. The ratio of the cavity’s axial gap width s to the disk radius r is a characteristic dimension describing the cavity, referred to as non-dimensional axial gap width G. G = (1) Between the disk vibration modes and the acoustic modes inside the ﬂuid-ﬁlled cavities, ﬂuid-structure interaction occurs if there is a resulting force from one to the other. These interactions present themselves in the form of the out-of-plane structural modes of the disk and ﬂuctuating pressure patterns of the ﬂuid that can only truly be neglected if the disk rotates in a vacuum. Structural and acoustic modes additionally inﬂuence each other if a certain mode pair is coupled. This results in a further frequency shift. In the case of coupling, the modes are referred to as structure and acoustic dominant modes. 3.2. Uncoupled Structure/Disk Vibrational Modes The mode shape of disk vibration modes, also called structural modes, is characterized by the number of nodal diameters (diametral modes, or circumferential order) m and the number of nodal circles (or radial order) n. For a rotating disk, a resonance between a natural frequency f of one of its disk modes and an excitation frequency f occurs when m,n e,st the latter coincides with the rotational frequency of the forward or backward traveling wave of the disk mode as follows [3]: f = f m f (2) e,st m,n W Int. J. Turbomach. Propuls. Power 2022, 7, 15 4 of 22 when the latter coincides with the rotational frequency of the forward or backward trav- eling wave of the disk mode as follows [3]: Int. J. Turbomach. Propuls. Power 2022, 7, 15 4 of 21 𝑓 = 𝑓 ±𝑚 × 𝑓 , , Ω (2) with 𝑓 being the rotational frequency of the rotating disk. with f being the rotational frequency of the rotating disk. Disk natural frequencies 𝑓 increase slightly with an increase in disk rotational Disk natural frequencies f increase slightly with an increase in disk rotational speed Ω due to the stiffening em ffec ,n t of the centrifugal force field [7]. This effect can be speed W due to the stiffening effect of the centrifugal force ﬁeld [7]. This effect can be approximated (as, e.g., in [2]) but is often sufficiently small to be neglected [7]. approximated (as, e.g., in [2]) but is often sufﬁciently small to be neglected [7]. For a mistuned disk, i.e., a non-axisymmetric disk, there are two natural frequencies For a mistuned disk, i.e., a non-axisymmetric disk, there are two natural frequencies for one disk mode. However, this frequency splitting is often negligible for practical ap- for one disk mode. However, this frequency splitting is often negligible for practical plications [7]. In general, the above-mentioned considerations also apply to rotating im- applications [7]. In general, the above-mentioned considerations also apply to rotating pellers. Nevertheless, mistuning due to geometrical non-uniformities becomes more likely impellers. Nevertheless, mistuning due to geometrical non-uniformities becomes more for more complex impeller geometries [5]. Such detuning gives rise to more irregular and likely for more complex impeller geometries [5]. Such detuning gives rise to more irregular complex mode shapes and more resonances than those in equivalent tuned systems [5]. and complex mode shapes and more resonances than those in equivalent tuned systems [5]. 3.3. Uncoupled Acoustic/Cavity Modes 3.3. Uncoupled Acoustic/Cavity Modes In a fluid-filled cavity, acoustic modes (or cavity modes) can form that establish as In a ﬂuid-ﬁlled cavity, acoustic modes (or cavity modes) can form that establish as fluctuating pressure patterns. In rotating machinery, these acoustic waves are excited by ﬂuctuating pressure patterns. In rotating machinery, these acoustic waves are excited by the flow through the machine and can be enhanced by the impeller rotation [10]. Similar the ﬂow through the machine and can be enhanced by the impeller rotation [10]. Similar to to structural modes, acoustic modes are characterized by their number of nodal diameters structural modes, acoustic modes are characterized by their number of nodal diameters p, p, nodal circles q, and additionally, axial nodes r. The frequency of the (p,q) acoustic mode nodal circles q, and additionally, axial nodes r. The frequency of the (p,q) acoustic mode inside a cylinder can be calculated as follows [24]: inside a cylinder can be calculated as follows [24]: ( ) 𝜋𝛼 𝑐 𝑓 = (3) , (pa) c p,q 2𝜋 𝑟 f = (3) p,q 2p r where (𝜋𝛼 ) are the dimensionless solutions to the first derivative of the Bessel function for the (p,q) acoustic mode (see Figure 1). From Equation (3) we see that the natural fre- where (pa) are the dimensionless solutions to the ﬁrst derivative of the Bessel function p,q for quenc they o (pf ,q the ) acoustic acoustic mode mode (see is line Figur arly e dependent o 1). From Equation n the fluid’ (3) we s speed o see that f sothe und natural 𝑐 , and fr thus on the equency of fluid proper the acoustic tie mode s. E.g., is an inc linearly rease dependent in fluid pressure causes a on the ﬂuid’s speed shift of in th sound e natur c, and al thus on the ﬂuid properties. E.g., an increase in ﬂuid pressure causes a shift in the natural frequency of the acoustic modes. This will be referred to as the speed-of-sound effect in frequency of the acoustic modes. This will be referred to as the speed-of-sound effect in the following. A change in cavity radius 𝑟 also results in a change of acoustic frequency; the following. A change in cavity radius r also results in a change of acoustic frequency; however, according to this model in Equation (3), a change in axial gap width does not. however, according to this model in Equation (3), a change in axial gap width does not. Figure 1. Non-dimensional solutions to Bessel functions for acoustic frequencies inside a cylinder Figure 1. Non-dimensional solutions to Bessel functions for acoustic frequencies inside a cylinder and respective mode shapes for different numbers of nodal diameters and nodal circles. and respective mode shapes for different numbers of nodal diameters and nodal circles. Int. J. Turbomach. Propuls. Power 2022, 7, 15 5 of 21 In a rotor-stator system, disk rotation results in ﬂuid rotation inside the cavity. These ﬂow patterns inﬂuence the formation of acoustic modes and can be described depending on the circumferential Reynolds number Re, which is deﬁned as follows: u L 2p W r Re = = (4) n n where u is the circumferential velocity, L the characteristic length, and n the kinematic viscosity of the ﬂuid. The widely used rotor-stator cavity ﬂow model by Daily and Nece (1960) [25] distinguishes between four ﬂow regimes, the two turbulent ones having either merged or separated boundary layers. The latter is predominant for bigger cavity widths or higher Reynolds numbers and is characterized by a core ﬂow, for which an almost constant circumferential velocity is assumed, i.e., the ﬂuid is swirling as a bulk ﬂow. The core rotation can be described by the ratio k of the ﬂuid velocity v to the impeller 0 f circumferential velocity u and only depends on the geometry of the cavity if there is no superimposed through ﬂow [26]. In the case of zero leakage ﬂow, it is also independent of the radius and can be estimated as follows [26]: k = = (5) r r t f ,case c c ax 1 + + 5 r r r c d d d f ,d where r and r are the impeller and the cavity radius, t is the cylindrical portion of c ax the impeller side room, and c and c are the friction coefﬁcients of the impeller and f ,d f ,case casing, respectively. Typically, the core rotation factor takes a value between 0.4 and 0.5, i.e., the ﬂuid rotates at a rotational speed a little less than half of the disk speed. Analogous to the structural modes, the acoustic modes subject to core rotation form a backward and a forward traveling wave in the cavities that can be calculated as follows: f = f k p f (6) e,ac p,q 0 W 3.4. Fluid-Structure Interaction So far, the disk vibration and ﬂuid-ﬁlled cavity have been treated separately. However, the inﬂuence of the surrounding ﬂuid on the impeller has to be considered. Especially with an increase in ﬂuid pressure and therefore density, signiﬁcant forces are acting from the ﬂuid on the impeller. Vice-versa, the out-of-plane structural modes also inﬂuence the acoustic ones. These effects are called ﬂuid-structure-interaction and cannot be neglected. Of particular practical interest is the inﬂuence of an increased ﬂuid density on the turboma- chinery impeller, which is addressed by the added-mass-effect: An increase in mass of the surrounding ﬂuid results in a decrease of the natural frequencies of the impeller [12,13,27], as the ﬂuid mass is linearly dependent on the ﬂuid density (and the cavity volume). Hence, while the impeller ’s mass does not change, the total mass of the system changes as the ﬂuid mass acting on the impeller varies. The change in natural frequency is anti-proportional to the change in total mass. However, these interactions are not of major concern unless coupling between a pair of acoustic and structural modes exists. 3.5. Coupling The respective modes need to be excited ﬁrst off for coupling between an acoustic and a structure mode to occur. While acoustic modes are excited by the ﬂuid ﬂow through the casing passing obstacles, such as vanes or the rotor, the most efﬁcient source of impeller vibration are higher order acoustic modes [10]. In addition, for a setup with two cavities surrounding an impeller or disk, there must be a phase shift between the acoustic modes in these cavities, or in a different notion, an axial node must exist (at the axial position of the disk). Only then is there a resulting force from the ﬂuid acting on the impeller. Therefore, Int. J. Turbomach. Propuls. Power 2022, 7, 15 6 of 22 Int. J. Turbomach. Propuls. Power 2022, 7, 15 6 of 21 Therefore, the out-of-plane structural (disk) modes and acoustic (cavity) modes with an the out-of-plane structural (disk) modes and acoustic (cavity) modes with an axial node axial node have the most potential to influence each other. have the most potential to inﬂuence each other. The compatibility of mode shapes is a necessary condition for structural-acoustic The compatibility of mode shapes is a necessary condition for structural-acoustic coupling. This coincidence of mode shapes is especially established through the same coupling. This coincidence of mode shapes is especially established through the same number of nodal diameters (m = p) [10]. It is often assumed that each structural mode is number of nodal diameters (m = p) [10]. It is often assumed that each structural mode is only well coupled to one acoustic mode and only weakly to all others [10,28]. The most only well coupled to one acoustic mode and only weakly to all others [10,28]. The most severe interaction (or complete coincidence) occurs when also the (uncoupled) natural severe interaction (or complete coincidence) occurs when also the (uncoupled) natural frequencies match [10]. However, any coupling of a pair of acoustic and structure domi- frequencies match [10]. However, any coupling of a pair of acoustic and structure dominant nant modes results in the shift of their natural frequencies compared to the uncoupled modes results in the shift of their natural frequencies compared to the uncoupled modes. modes. This coupling effect also occurs if the theoretically uncoupled natural frequencies This coupling effect also occurs if the theoretically uncoupled natural frequencies are in a are in a certain proximity to each other. The closer the uncoupled natural frequencies are certain proximity to each other. The closer the uncoupled natural frequencies are to each to each other, the stronger is the coupling effect and thus the resulting frequency shift. other, the stronger is the coupling effect and thus the resulting frequency shift. According According to [10], a 25% minimum separation of natural frequencies is recommended to to [10], a 25% minimum separation of natural frequencies is recommended to consider consider modes as not coupled. In Reference [11], the degree of structural-to-acoustic cou- modes as not coupled. In Reference [11], the degree of structural-to-acoustic coupling is pling is expressed by a coupling coefficient, which is zero in the case of uncoupled sys- expressed by a coupling coefﬁcient, which is zero in the case of uncoupled systems. A tems. A small mass of the impeller or a thin cavity increases the coupling, while a bigger small mass of the impeller or a thin cavity increases the coupling, while a bigger mass of mass of the disk and a wider cylindrical cavity facilitate weaker coupling [11]. the disk and a wider cylindrical cavity facilitate weaker coupling [11]. Still, it should be mentioned that for real turbomachinery impellers or detuned Still, it should be mentioned that for real turbomachinery impellers or detuned bladed bladed disk assemblies, each excited disk mode contains components of many diametral disk assemblies, each excited disk mode contains components of many diametral mode mode orders. Therefore, an m-engine order excitation might excite not only the pure m- orders. Therefore, an m-engine order excitation might excite not only the pure m-nodal nodal diameter mode but also most other modes [5]. diameter mode but also most other modes [5]. 3.6. Existing Frequency Shift Models 3.6. Existing Frequency Shift Models One wide One widely ly used mathem used mathematical atical model aiming model aiming at at explaining couplin explaining coupling g effec effects ts in ro in rtor otor - - ca cavity vity syst systems ems is is th the e model of Ma model of Magara gara et et aal. l. (200 (2008) 8) [12] [12, whi ], which ch origi originates nates frfr om om resea resear rch on ch on freq frequency uency sh shifts ifts of of cen centrifugal trifugal compres compressor sor impel impellers lers in in high high-density -density ggas as appl applications. ications. The frequency shift model consists of a parabola, which opens upward (left side of The frequency shift model consists of a parabola, which opens upward (left side of Eq Equation uation (7 (7)) )) aand nd aa strai straight ght liline ne wit with h a aposit positive ive slslope ope (r(right ight siside de of of Eq Equation uation (7 (7)) )) ifif th the e square of the system frequency f is indicated on the x-axis (see Figure 2). square of the system frequency 𝑓 is indicated on the x-axis (see Figure 2). 𝜌 𝑐 2 r c 𝑓 2 − 𝑓 2 ∗ 𝑓 2 − 𝑓 2= 2 𝑓 2 (7) , , f f f f = 2 f (7) 1,uc s 2,uc s 𝜌 𝑏 𝑠 s r b s d c Figure 2. Coupling model of Magara et al. (2008) [12]. Figure 2. Coupling model of Magara et al. (2008) [12]. The intersections of the parabola with the x-axis (points A) represent the theoretically The intersections of the parabola with the x-axis (points A) represent the theoretically uncoupled natural frequencies of the disk and the fluid. Which of them is the lower and uncoupled natural frequencies of the disk and the ﬂuid. Which of them is the lower and which whichthe h the higher igher-frequenc -frequency y mode mode depends depends on on the the ge geometry ometryand andcannot cannot be stated gene be stated generally r- . ally. There Therefore,fthey ore, they are not are no indicated t indic as ated acoustic as aco and ustic st an ructur d struc e modes, ture mo butdes, but as f and as 𝑓 f and . The 1,uc , 2,uc 𝑓 frequency . The fre dif quen ference cy differenc between e between A and A 𝐴 indicates and 𝐴 the ind coupling icates the couplin potentialg p of the otent two ial o m fodes, the 1 2 two mo being bigger des, bethe ing big closer gerthe the clo theor ser etically the theoretically uncoupleduncoupled natural frequencies natural fre ar qe uenc to each ies are other to . each oth The straight er. Thline e strindicates aight line indic the coupling ates the str couplin ength.g streng The slope th. The increases slope incr witheases with an an increase in ﬂuid density r and speed of sound c and decreases with an increase in disk density r , increase in fluid density 𝜌 and speed of sound c and decreases with an increase in disk f d Int. J. Turbomach. Propuls. Power 2022, 7, 15 7 of 22 Int. J. Turbomach. Propuls. Power 2022, 7, 15 7 of 21 with an increase in fluid density 𝝆 and speed of sound c and decreases with an increase in disk density 𝝆 , disk thickness 𝒃 , and axial cavity width 𝒔 . The intersections of the 𝒅 𝒄 coupling strength term with the parabola (points B and C) represent the two natural disk thickness b, and axial cavity width s . The intersections of the coupling strength term frequencies of the coupled system 𝒇 . When the coupling strength increases, the with the parabola (points B and C) represent the two natural frequencies of the coupled intersections are shifted upwards and outwards, i.e., the frequency of the lower-frequency system f . When the coupling strength increases, the intersections are shifted upwards and mode decreases (from 𝑩 to 𝑪 ) and the one of the higher-frequency mode increases 𝟏 𝟏 outwards, i.e., the frequency of the lower-frequency mode decreases (from B to C ) and 1 1 (from 𝑩 to 𝑪 ). This frequency shift is bigger, the more the coupling strength increases 𝟐 𝟐 the one of the higher-frequency mode increases (from B to C ). This frequency shift is 2 2 (the bigger its slope), and the closer the theoretically uncoupled modes are to each other, bigger, the more the coupling strength increases (the bigger its slope), and the closer the i.e., the bigger the coupling potential. The latter effect can be illustrated by comparing the theoretically uncoupled modes are to each other, i.e., the bigger the coupling potential. The frequency shifts of the cases with weak (solid line parabola) and strong coupling potential latter effect can be illustrated by comparing the frequency shifts of the cases with weak (dashed line parabola). (solid line parabola) and strong coupling potential (dashed line parabola). 3.7. 3.7.Overview Overviewon on Fr Freq equency uency Shift Shift Phenomena Phenomena in a Rotor-Cavity-System in a Rotor-Cavity-System Based Based on on t the he com compr preh ehensive ensive review review of of struc structural tural and and acou acoustic stic momodes, des, as we asll well as thas eir their ﬂuid-structure interaction and coupling in the previous paragraphs, an overview of fluid-structure interaction and coupling in the previous paragraphs, an overview of frequency shift effects in rotor-cavity systems is presented in what follows (Figure 3). frequency shift effects in rotor-cavity systems is presented in what follows (Figure 3). Figure 3. Overview of frequency shift effects in rotor-cavity systems [12]. Figure 3. Overview of frequency shift effects in rotor-cavity systems [12]. First, First, ther ther ee is is aa gr gro oup up of of “uncoupled “uncoupled mode mode ef ef fects”, fects”,which which occur occur independently independentlyfr from om any coupling of a speciﬁc mode pair, i.e., they can be observed for both, uncoupled as well as any coupling of a specific mode pair, i.e., they can be observed for both, uncoupled as well for coupled modes: The speed of sound effect refers to a shift in the natural frequency of the as for coupled modes: The speed of sound effect refers to a shift in the natural frequency acoustic mode if the speed of sound changes due to altered ﬂuid properties. Additionally, of the acoustic mode if the speed of sound changes due to altered fluid properties. the stiffening effect of the centrifugal force ﬁeld might be observed for rotating impellers Additionally, the stiffening effect of the centrifugal force field might be observed for resulting in a slight upward shift of the structural frequency at higher rotational speeds. rotating impellers resulting in a slight upward shift of the structural frequency at higher Considering ﬂuid-structure interaction effects, which are of particular importance for higher rotational speeds. Considering fluid-structure interaction effects, which are of particular ﬂuid densities surrounding an impeller, the added mass effect plays an important role. importance for higher fluid densities surrounding an impeller, the added mass effect It results in a decrease of disk frequencies if the mass of the surrounding ﬂuid increases. plays an important role. It results in a decrease of disk frequencies if the mass of the In Magara’s coupling model [12], these effects are encompassed by the term “coupling surrounding fluid increases. In Magara’s coupling model [12], these effects are strength”. If two speciﬁc modes are coupled, i.e., a structure and an acoustic dominant encompassed by the term “coupling strength”. If two specific modes are coupled, i.e., a mode with similar mode shapes and natural frequencies in a certain proximity to each structure and an acoustic dominant mode with similar mode shapes and natural other are excited, this adds additional frequency shifts: For a coupled mode pair, one frequencies in a certain proximity to each other are excited, this adds additional frequency natural frequency shifts towards higher, the other towards lower frequencies, compared shifts: For a coupled mode pair, one natural frequency shifts towards higher, the other to their uncoupled counterparts. This is referred to as the coupling effect. The smaller the towards lower frequencies, compared to their uncoupled counterparts. This is referred to frequency difference of the uncoupled modes is, the bigger is the coupling effect, and thus as the coupling effect. The smaller the frequency difference of the uncoupled modes is, the frequency shift it effectuates. In Magara’s coupling model [12], these frequency shifts the bigger is the coupling effect, and thus the frequency shift it effectuates. In Magara’s are addressed by the term “coupling potential”. coupling model [12], these frequency shifts are addressed by the term “coupling Based on this, the following can be hypothesized: Any frequency shift that is not potential”. caused by the added mass, speed of sound, or stiffening effect (for uncoupled structural or acoustic modes) indicates an additional coupling effect. The coupled modes are then referred to as structure dominant and acoustic dominant modes. Int. J. Turbomach. Propuls. Power 2022, 7, 15 8 of 22 Int. J. Turbomach. Propuls. Power 2022, 7, 15 8 of 21 or acoustic modes) indicates an additional coupling effect. The coupled modes are then referred to as structure dominant and acoustic dominant modes. When the frequency difference between two coupled modes decreases subject to, e.g., When the frequency difference between two coupled modes decreases subject to, a change in fluid properties or rotational speed, the two modes approach each other. How- e.g., a change in ﬂuid properties or rotational speed, the two modes approach each other. ever, in several cases, they do not intersect (no resonance) but diverge again and approach However, in several cases, they do not intersect (no resonance) but diverge again and the theoretically uncoupled frequency path of the other mode. This frequency deflection approach the theoretically uncoupled frequency path of the other mode. This frequency is called mode veering and reported, e.g., in [27,29]. deﬂection is called mode veering and reported, e.g., in [27,29]. 4. Experimental Setup and Procedure 4. Experimental Setup and Procedure 4.1. Test Rig and Instrumentation 4.1. Test Rig and Instrumentation Experimental results presented in this paper are obtained at the rotor-cavity test rig Experimental results presented in this paper are obtained at the rotor-cavity test rig of the Chair of Turbomachinery at the University Duisburg-Essen, Germany, introduced of the Chair of Turbomachinery at the University Duisburg-Essen, Germany, introduced in [21–23]. The test rig consists of a plain rotating disk with a constant width 𝑏 in a cylin- in [21–23]. The test rig consists of a plain rotating disk with a constant width b in a cylindri- drical stationary casing, forming two fluid-filled cavities, which are connected through a cal stationary casing, forming two ﬂuid-ﬁlled cavities, which are connected through a small small radial gap. A sketch of the test rig, including some measurement equipment, is given radial gap. A sketch of the test rig, including some measurement equipment, is given in in Figure 4. This setup represents a simplified model of the side cavities in radial com- Figure 4. This setup represents a simpliﬁed model of the side cavities in radial compressors. pressors. The axial gap widths of the front and rear cavity, 𝑠 and 𝑠 , respectively, can The axial gap widths of the front and rear cavity, s and s , respectively, can be adjusted c f cr be adjusted independently. For the conducted experiments, there is no superimposed independently. For the conducted experiments, there is no superimposed through ﬂow through flow in the cavities. The disk is mounted on a shaft connected to a continuously in the cavities. The disk is mounted on a shaft connected to a continuously adjustable adjustable high-speed electrical drive via a magnetic coupling. Different gases can be high-speed electrical drive via a magnetic coupling. Different gases can be passed into the passed into the test rig from a pre-mix chamber, allowing for a variation of fluid parame- test rig from a pre-mix chamber, allowing for a variation of ﬂuid parameters. The disk’s ters. The disk’s rotational speed 𝛺 can be varied, and circumferential Reynolds numbers rotational speed W can be varied, and circumferential Reynolds numbers up to 2 108 can up to 2 × 108 can be achieved with CO2 at maximum speed. be achieved with CO at maximum speed. Figure 4. Test rig setup and measurement equipment. Figure 4. Test rig setup and measurement equipment. Excitation sources in real turbomachinery, such as rotor-stator interaction, are miss- Excitation sources in real turbomachinery, such as rotor-stator interaction, are missing ing in the test rig due to its simplified (blade- and vane-less) geometry. For that reason, a in the test rig due to its simpliﬁed (blade- and vane-less) geometry. For that reason, a loudspeaker is installed in the casing of the front cavity. It is capable of exciting both loudspeaker is installed in the casing of the front cavity. It is capable of exciting both a acoustic coustic a and nd s str tru uctur cture e d dominant ominant modes modes iin n t the he fr front ont a and nd the the r rear ear c cavities. avities. The The e excitation xcitation frequency is set by a wave generator connected to an ampliﬁer and adjusted independently from other parameters. The geometry of the test rig is designed so that at least one pair of Int. J. Turbomach. Propuls. Power 2022, 7, 15 9 of 21 acoustic and structural modes with the same number of nodal diameters are coupled at ambient conditions. When interpreting the results, one should notice that the cavity, in which the mea- surements are taken, is not perfectly closed. There are two small openings in the front cavity, one at the outer radius and one at the inner. Although valves in the connecting pipes are closed so that there is no in- or outﬂow to or from the cavity, these openings might affect the propagation and reﬂection of acoustic waves inside the cavity. In future experiments, results for a completely closed cavity and the current setup will be compared to gain further insight into this matter. However, the side cavities in real turbomachines are also not perfectly closed, so that the setup in this work might even represent a more realistic case than that of a perfectly sealed cavity. To measure acoustic modes that occur in the form of local pressure ﬂuctuations, the test rig is equipped with two pressure transducers in each, the front and rear cavity, as well as one on the disk. To detect structural modes in the form of disk vibration, two strain gauges are mounted on the disk. The following measures are provided for the identiﬁcation of mode shapes: The number of nodal diameters of the excited mode is traced with an additional traversable pressure sensor installed in the front cavity, which can rotate 300 degrees around the disk’s circumference at a constant excitation frequency. To detect the number of nodal circles, the two pressure sensors in the front cavity are mounted at a different radial position compared to those in the rear cavity. An axial node is detected by identifying a 180 degrees’ phase shift between the pressure patterns in the front and rear cavity. In addition, the test rig and pre-mix chamber are equipped with pressure and temperature sensors to determine the thermodynamic properties of the ﬂuid. The general uncertainty of the measurement chain of the temperature sensors is determined to be 1.8 K, and the non-linearity and dynamic uncertainties of the pressure sensors are negligibly small [29]. However, strain gauges are only able to measure dynamic vibrations up to a cut-off frequency of 50 kHz [29]. Additionally, the electrostatic noise increases signiﬁcantly for experiments with a rotating disk, resulting in a very low signal- to-noise ratio for those strain gauge results. However, pressure sensor results are not affected as much by electromagnetic interferences. Therefore, response functions for the structure dominant modes are also measured with pressure sensors in case of disk rotation. Measurements without disk rotation have shown that this approach does not inﬂuence the results as strain gauge and pressure sensor measurements agree well. 4.2. Experimental and Evaluation Procedures In preparation for the experiments, the test rig is ﬁlled with the required gas mixture until the target pressure is reached. The temperature inside the test rig is assumed to be constant, as a maximum friction-induced increase of 0.3% for experiments with high rotational speeds is considered negligible. All experiments aiming to identify the modal parameters of the excited modes are conducted through frequency sweep experiments. First, the frequency interval and sweep velocity are set. The latter is chosen sufﬁciently small so that the transducer ’s response can be considered stationary. The frequency generator produces the respective harmonic signal, ampliﬁed to drive the loudspeaker acting as the excitation source. Sweep experiments are conducted with a constant sweep velocity from low to high frequencies and the rotatable pressure sensor set to a ﬁxed position. Each frequency sweep experiment is repeated three times. Additional measurements with the rotatable pressure sensor are conducted at a con- stant excitation frequency for mode order identiﬁcation. The rotatable pressure sensor traverses to different circumferential positions, where it remains for the duration, during which pressure ﬂuctuations are measured at this position. In this way, pressure patterns in circumferential direction are identiﬁed, and the number of nodal diameters is determined. Moreover, a coupled ﬁnite element method analysis and an experimental hammer test of Int. J. Turbomach. Propuls. Power 2022, 7, 15 10 of 21 the disk have been conducted to have a reference for the modal parameters of the disk vibration modes. The data from the pressure transducers and strain gauges recorded during the sweep experiments are evaluated by Fourier transformation. Thereby, the raw data is trans- formed from the time into the frequency domain, and frequency response functions are obtained for each sensor. Natural frequency results are obtained using the peak picking method. However, especially for acoustic dominant modes in the “slim cavity” setup, many modes overlap, inﬂuence, cross, and deﬂect each other. Moreover, interference signals inﬂuence the measured frequency response, so that frequency peaks cannot always be identiﬁed undoubtedly. 4.3. Conducted Experiments Experiments are conducted with two different geometrical setups. For both, the rear cavity has a non-dimensional axial gap width of 0.1. The front cavity width in the “wide cavity” setup is symmetrical to the rear cavity, while for the “slim cavity” it has a non- dimensional axial gap width of 0.025. As shown in the following, the front cavity is the “dominant” one, determining the formation of acoustic (dominant) modes, likely because the excitation source is located here. Different experimental series are conducted, aiming to reveal natural frequency shifts due to different effects discussed theoretically beforehand. In the ﬁrst test series, “GAP”, a brief comparison of results for different axial gap widths is presented. A change in geometry causes a change in coupling strength by altering the ﬂuid mass acting on the disk. In the second test series “PRESSURE”, the ﬂuid pressure in the cavity is changed, effectuating the coupling strength. Density and speed of sound of the ﬂuid change at the same time. Therefore, an added mass and speed of sound effects are expected. While the ﬁrst two experimental series are conducted with a non-rotating disk, the third one, “ROTATION”, includes experiments with different disk rotational speeds, changing the coupling potential between modes by altering the frequency difference between them. Through this, the coupling potential is expected to be signiﬁcantly increased, resulting in much stronger coupling effects than for the previous test series. An overview of the conducted experimental series is given in Table 1. Table 1. Overview of conducted test series and expected frequency shift effects. Expected Frequency Shift Altered Coupling Possible Further Frequency Test Series Description Effect(s) for Inﬂuence Shift Due to Coupling Uncoupled Modes “GAP” Variation of axial gap width Variation of coupling strength Added mass effect Weak coupling effect Added mass effect, speed of “PRESSURE” Variation of ﬂuid pressure Variation of coupling strength Weak coupling effect sound effect Stiffening effect of centrifugal “ROTATION” Variation of disk rotational speed Variation of coupling potential Coupling effect force ﬁeld Based on the preceding remarks, two main hypotheses are stipulated: 1. Any frequency shift that cannot be explained by “uncoupled mode effects” is presum- ably due to additional coupling effects; 2. The coupling effect is stronger the closer the natural frequencies of the theoretically uncoupled modes are to each other. To validate these hypotheses, the experimental results in this work are assessed subject to the following approach: 1. Theoretic frequency shifts of “uncoupled mode effects” are quantiﬁed and measured frequency shifts corrected for their inﬂuence; 2. The remaining frequency shifts are analyzed, which are expected to be due to “cou- pling effects”. Int. J. Turbomach. Propuls. Power 2022, 7, 15 11 of 21 4.4. Estimation of Coupling Effect To estimate the coupling effect, i.e., the extent of any frequency shift caused by cou- pling, a coupling potential parameter CP and a coupling strength parameter CS are intro- duced as follows: f f ac st CP = (8) min( f , f ) ac st r c CS = (9) c f The coupling potential parameter CP represents the frequency separation between the coupled modes. A smaller value indicates stronger coupling. It is positive when the acoustic dominant mode has a higher frequency than the structure dominant one, and negative if it is vice versa. In the following, the coupling potential is deﬁned as “strong” when the frequency separation between the structure and acoustic dominant modes is less than 0.3. The coupling strength parameter CS depends on the ﬂuid properties as well as the geometry of the cavity. Both coupling parameters together allow concluding on a possible coupling effect. 4.5. Presentation of Measurement Results Measurement results are presented in the following way: Instead of absolute natural frequencies f and f , normalized frequencies f are used in the following, which are st ac frequencies relative to that of the two nodal diameter structure mode in air for the “slim cavity” setup, which serves as a reference frequency f , as follows: re f f f f = = (10) f f air, 0.1 MPa, 295 K ( ) st,2,0 re f Normalized frequencies f can be calculated for any structure (dominant) mode (m,n) and any acoustic (dominant) mode (p,q,r). For the evaluation of frequency shifts, relative frequency changes D f are displayed. These are frequency changes relative to the frequency at a cavity pressure of 0.1 MPa as follows: f f f f 0.1 MPa ( ) D f = = (11) f f (0.1 MPa) Corrected relative frequency changes D f of structural modes (corrected for the st,corr added mass effect) are calculated as follows, with m = m + m being the total mass of tot d f disk and ﬂuid together: m (r) tot f (r) f st st,0 f (r) f m st,corr st,0 tot,0 D f (r) = = (12) st,corr f f st,0 st,0 Corrected relative frequency changes D f of acoustic modes (corrected for the ac,corr speed of sound effect) are calculated as follows: f (r) f ac ac,0 f (r) f c(r) ac,corr ac,0 D f (r) = = (13) ac,corr f f ac,0 ac,0 Corrected relative frequency changes D f and D f can be calculated for any st,corr ac,corr structure (dominant) mode (m,n) and any acoustic (dominant) mode (p,q,r), respectively. 5. Measurement Results and Discussion Acoustic and structure dominant modes that can be excited in the test rig within the 1 1 range of analyzed frequencies (150 s to 1400 s ) are the two to ﬁve nodal diameter modes with zero nodal circles. Following the theory of a match of mode shapes [10], only modes with the same mode order are considered as potentially coupled (m = p, n = q). Int. J. Turbomach. Propuls. Power 2022, 7, 15 12 of 21 Coupled acoustic dominant modes have an additional axial node (r = 1), so that there is a net force acting on the disk. For all analyzed mode pairs, except for the ﬁve nodal diameters one in the “slim cavity”, the acoustic dominant mode has a higher frequency than the structure dominant mode. This, along with other characteristic data of the occurring mode pairs, is displayed in Table 2. Table 2. Overview of excited mode couples (with n = q = 0) in test rig at UDE. The coupling strength is calculated for 0.1 MPa and 295 K. Values of the coupling strength parameter are normalized as a ratio of the respective value to that for the case of air in the “slim cavity”. Nodal Estimated Test Rig Setup k Fluid CS/(CS ) f f CP 0 slim,air st ac Diameters (m = p) Coupling Effect 3 1.99 8.86 * +3.45 uncoupled helium 0.101 4 3.44 10.75 * +2.13 uncoupled Wide Cavity 0.44 2 1.01 2.41 +1.37 weak 3 1.99 3.02 +0.52 weak air 0.250 4 3.40 3.66 +0.08 (very) strong 2 1.00 2.53 +1.53 weak 3 1.98 3.78 +0.90 weak Slim Cavity 0.48 air 1 4 3.42 4.38 +0.28 strong 5 5.26 4.61 0.14 strong * Frequencies of the acoustic modes in helium are outside of the measured frequency interval, and therefore estimated based on ﬂuid properties and ﬁrst derivative of Bessel function. The different geometries of the two test rig setups result in slightly different core rotation factors k and thereby different coupling strength parameters CS. The latter is also controlled by the ﬂuid properties. The frequency separation between the theoreti- cally uncoupled structure and acoustic modes with the same number of nodal diameters determines the coupling potential parameter CP. Both coupling parameters then predict the coupling effect. In Table 2, coupling strength parameters CS are exemplarily given for 0.1 MPa and 295 K, and coupling potential parameters CP for standstill of the disk. From this preliminary estimation of the coupling effect, the coupling strength is expected to be bigger for the “slim cavity” test rig setup compared to the “wide” one; and lower in Helium compared to air. The coupling potential is estimated to be the biggest for the four nodal diameter mode in air for the “wide cavity” setup, followed by the ﬁve and four nodal diameter modes in the “slim cavity”. For these mode pairs, frequency shifts are expected to be observable due to the strong coupling potential. Results of the three different test series (as introduced in Table 1) are presented in the following. 5.1. Variation of Axial Gap Width In the ﬁrst experimental series, “GAP”, the axial gap width of the front cavity is varied and thereby is the coupling strength. The disk is non-rotating. With a decrease in (front) cavity width, the ﬂuid mass acting on the disk decreases, and the coupling strength increases (see Table 2 for CS values). Response functions recorded for the different front cavity widths differ considerably from each other. With a decrease in the cavity width, the natural frequencies of the structure dominant modes change only slightly (max 2%), while the acoustic dominant modes shift signiﬁcantly towards higher frequencies (5% to 25%). Therefore, it can be concluded that the axial gap width has a signiﬁcant inﬂuence on the occurring acoustic dominant modes. Additionally, more peaks appear to occur in the response functions, and many peaks overlap in the measurements with a smaller front cavity width. The measured natural frequencies do not reﬂect the theoretically expected behavior if no coupling is anticipated (slight increase in disk natural frequencies and no change in acoustic dominant modes), neither do they exhibit the expected frequency shifts if coupling is assumed. Consequently, neither the “uncoupled mode effects” nor the coupling models Int. J. Turbomach. Propuls. Power 2022, 7, 15 13 of 22 that the axial gap width has a significant influence on the occurring acoustic dominant modes. Additionally, more peaks appear to occur in the response functions, and many peaks overlap in the measurements with a smaller front cavity width. The measured natural frequencies do not reflect the theoretically expected behavior if no coupling is anticipated (slight increase in disk natural frequencies and no change in acoustic dominant modes), neither do they exhibit the expected frequency shifts if cou- Int. J. Turbomach. Propuls. Power 2022, 7, 15 13 of 21 pling is assumed. Consequently, neither the “uncoupled mode effects” nor the coupling models can explain the observed frequency shifts subject to a variation of the axial gap width, i.e., the theory does not explain the measurement results. can explain the observed frequency shifts subject to a variation of the axial gap width, i.e., the theory does not explain the measurement results. 5.2. Presure Variation In the second experimental series “PRESSURE”, with a non-rotating disk, the fluid 5.2. Presure Variation pressure inside the test rig is changed while all other parameters remain constant. In the second experimental series “PRESSURE”, with a non-rotating disk, the ﬂuid Thereby, the coupling strength is varied (for CS values at 0.1MPa see Table 2), and added pressure inside the test rig is changed while all other parameters remain constant. Thereby, mass as well as speed of sound effects are expected to influence measured natural fre- the coupling strength is varied (for CS values at 0.1MPa see Table 2), and added mass as quencies. In addition, a weak coupling effect might be observable, depending on the ex- well as speed of sound effects are expected to inﬂuence measured natural frequencies. In cited mode pair (see CP values in Table 2). For all measurement results of the series “PRES- addition, a weak coupling effect might be observable, depending on the excited mode pair SURE”, corrected relative frequency changes ∆𝑓 are displayed as defined in Equations (see CP values in Table 2). For all measurement results of the series “PRESSURE”, corrected (11)–(13). relative frequency changes D f are displayed as deﬁned in Equations (11)–(13). corr A measurement series using the “wide cavity” setup and helium as the fluid is con- A measurement series using the “wide cavity” setup and helium as the ﬂuid is con- ducted as disk vibration modes measured in helium can be considered approximately un- ducted as disk vibration modes measured in helium can be considered approximately coupled due to the big frequency separation to the respective acoustic modes (see Table uncoupled due to the big frequency separation to the respective acoustic modes (see 2). If no further coupling is assumed, the frequencies should not change subject to a Table 2). If no further coupling is assumed, the frequencies should not change subject to change in fluid density after being corrected for the added mass effect. As shown in Figure a change in ﬂuid density after being corrected for the added mass effect. As shown in 5, natural frequencies of the measured three and four nodal diameter structure modes Figure 5, natural frequencies of the measured three and four nodal diameter structure decrease slightly with an increase in fluid density. Corrected for the added mass effect, an modes decrease slightly with an increase in ﬂuid density. Corrected for the added mass insignificant decrease of 0.3% and 0.2% remains for the (3,0) and (4,0) structural modes, effect, an insigniﬁcant decrease of 0.3% and 0.2% remains for the (3,0) and (4,0) structural respectively, if the fluid density is increased approximately 15-fold. This remaining de- modes, respectively, if the ﬂuid density is increased approximately 15-fold. This remaining crease in frequency is considered as negligible, and therefore these modes in helium can decrease in frequency is considered as negligible, and therefore these modes in helium can be be consid consider ere eddas as practically practically constant constantand and thus thus uncoupled. uncoupled. Figure 5. Corrected relative frequency changes of structure modes in helium for “wide cavity” over Figure 5. Corrected relative frequency changes of structure modes in helium for “wide cavity” over fluid density. Raw values and values corrected for added mass effect. ﬂuid density. Raw values and values corrected for added mass effect. Measurements using the “slim cavity” setup and air as the fluid are conducted to Measurements using the “slim cavity” setup and air as the ﬂuid are conducted to analyze analyze the the e excited xcited modes for potential couplin modes for potential coupling. g. As As mode pairs mode pairs in in air air exhibit exhibit aa higher higher coupling potential than those in helium, frequency shifts attributed to coupling effects are expected to be bigger in air, but dependent on the coupling potential of each mode pair. Corrected relative frequency changes of the structure dominant modes D f are st,corr presented in Figure 6. For the modes with two to four nodal diameters, the natural frequencies of the disk decrease about 4% to 5% for a 12-fold increase in ﬂuid density, while the frequency of the ﬁve nodal diameter structural mode increases about 5%. These frequency changes are considerably bigger than those measured in the almost uncoupled case for helium, and it can be concluded that there must be an additional coupling effect. Measured frequency shifts agree qualitatively with the coupling model of Magara et al. Int. J. Turbomach. Propuls. Power 2022, 7, 15 14 of 22 coupling potential than those in helium, frequency shifts attributed to coupling effects are expected to be bigger in air, but dependent on the coupling potential of each mode pair. Corrected relative frequency changes of the structure dominant modes ∆𝑓 are pre- sented in Figure 6. For the modes with two to four nodal diameters, the natural frequen- cies of the disk decrease about 4% to 5% for a 12-fold increase in fluid density, while the frequency of the five nodal diameter structural mode increases about 5%. These frequency Int. J. Turbomach. Propuls. Power 2022, 7, 15 14 of 21 changes are considerably bigger than those measured in the almost uncoupled case for helium, and it can be concluded that there must be an additional coupling effect. Meas- ured frequency shifts agree qualitatively with the coupling model of Magara et al. (2008) [12], indicating a decrease in the frequency of the structure dominant modes in case of (2008) [12], indicating a decrease in the frequency of the structure dominant modes in 𝑓 𝑓 , and an increase for the opposite case (for the five nodal diameter mode). How- case of f < f , and an increase for the opposite case (for the ﬁve nodal diameter mode). st ac ever, the frequency changes appear to be very similar for the two, three and four nodal However, the frequency changes appear to be very similar for the two, three and four nodal diameter diametermodes, modes, although although those those h have ave d dififer ffer ent ent coupling coupling poten potentials. tials. Figure Figure 6. 6. Corr Corrected ected r relati elative ve fr frequency equencychanges changes of of str stru uctur ctue re dominant dominant m modes odes in in air air for for “ “slim slimcavity” cavity” over fluid density. Values corrected for added mass effect. over ﬂuid density. Values corrected for added mass effect. While most frequency peaks are well visible for structural modes, this is not the While most frequency peaks are well visible for structural modes, this is not the case case for most aco for mostu acoustic stic modes; modes; the two the nodal d two nodal iame diameter ter mode is mode the is on the ly only exception. exception. To show To show potential coupling inﬂuences, both (acoustic and structure dominant) modes of the potential coupling influences, both (acoustic and structure dominant) modes of the two two nodal diameter mode pair are displayed in Figure 7. After the structural mode is nodal diameter mode pair are displayed in Figure 7. After the structural mode is corrected corrected for the added mass effect and the acoustic mode for the speed of sound effect, for the added mass effect and the acoustic mode for the speed of sound effect, both still both still exhibit a decreasing and a slightly increasing trend, respectively, when plotted exhibit a decreasing and a slightly increasing trend, respectively, when plotted over the over the ﬂuid density. This behavior supports the hypothesis that weak coupling is present fluid density. This behavior supports the hypothesis that weak coupling is present be- Int. J. Turbomach. Propuls. Power 2022, 7, 15 15 of 22 between the disk and the surrounding ﬂuid, and the trends agree qualitatively with those tween the disk and the surrounding fluid, and the trends agree qualitatively with those in in Magara’s [12] coupling model. Magara’s [12] coupling model. Figure 7. Corrected relative frequency changes of two nodal diameter mode pair in air for “slim Figure 7. Corrected relative frequency changes of two nodal diameter mode pair in air for “slim cav- cavity” over fluid density. Values corrected for added mass effect and speed of sound effect, respec- ity” over ﬂuid density. Values corrected for added mass effect and speed of sound effect, respectively. tively. Sensors: S8 on disk, S9 in front cavity. Sensors: S8 on disk, S9 in front cavity. It should be noted that the frequency change of the acoustic dominant mode differs It should be noted that the frequency change of the acoustic dominant mode differs slightly depending on the pressure sensor’s location in the test rig. It is smaller (+2.2% (S9) slightly depending on the pressure sensor ’s location in the test rig. It is smaller (+2.2% and +1.3% (S8 on disk)) than that of the structure dominant mode (−5.5%) for the approx- imately 14-fold increase in density. Altogether, for the measurement series “PRESSURE”, frequency shifts of structure dominant modes are observed that exceed the “uncoupled mode effects” and thus indi- cate a coupling effect, even for modes that are only weakly coupled. However, no clear picture about frequency shifts of acoustic dominant modes is obtained, and the hypothesis of a more pronounced coupling effect for stronger coupled modes could not be confirmed on the base of the measurement series with a variation in fluid pressure. 5.3. Variation of Disk Rotational Speed In the third experimental series, “ROTATION”, the disk rotational speed is varied, while the fluid pressure (0.1 MPa, air) and all other parameters remain constant. By this means, the coupling potential is changed as frequency differences between potentially coupled acoustic and structure dominant modes are varied (rotating case 𝐶𝑃 values dif- fer from standstill values in Table 2). There is no added mass or speed of sound effect. However, a small effect of the stiffening effect of the centrifugal force field of the impeller might be observable, but measurement values are not explicitly corrected for its influence. Disk rotation results in the formation of two modes for each mode order, one forward and one backward traveling wave. While disk modes theoretically travel at the disk rota- tional speed, cavity modes travel at the fluid’s rotational speed in the cavity. Assuming bulk flow in the cavity at approximately half the disk’s rotational speed, frequencies of acoustic and structure dominant modes approach each other subject to a change in disk rotational speed. These effects are shown in the following, first for structure, then for acoustic dominant modes, and last for specific mode pairs. For all measurement results of the series “ROTATION”, normalized frequencies 𝑓 are displayed as defined in Equation (10). In addition to the measurement results, theoretical values for the natural frequencies of structure and acoustic dominant modes at the respective disk rotational speed are given, calculated according to Equations (2) and (6), respectively. Utilized core rotation factors for the acoustic dominant modes are listed in Table 2. As shown in Figure 8, the development of a forward and a backward traveling wave is clearly visible for each mode, and the measured frequency splits of the disk vibration modes subject to rotation agree well with the theoretical values. Int. J. Turbomach. Propuls. Power 2022, 7, 15 15 of 21 (S9) and +1.3% (S8 on disk)) than that of the structure dominant mode (5.5%) for the approximately 14-fold increase in density. Altogether, for the measurement series “PRESSURE”, frequency shifts of structure dominant modes are observed that exceed the “uncoupled mode effects” and thus indicate a coupling effect, even for modes that are only weakly coupled. However, no clear picture about frequency shifts of acoustic dominant modes is obtained, and the hypothesis of a more pronounced coupling effect for stronger coupled modes could not be conﬁrmed on the base of the measurement series with a variation in ﬂuid pressure. 5.3. Variation of Disk Rotational Speed In the third experimental series, “ROTATION”, the disk rotational speed is varied, while the ﬂuid pressure (0.1 MPa, air) and all other parameters remain constant. By this means, the coupling potential is changed as frequency differences between potentially coupled acoustic and structure dominant modes are varied (rotating case CP values differ from standstill values in Table 2). There is no added mass or speed of sound effect. However, a small effect of the stiffening effect of the centrifugal force ﬁeld of the impeller might be observable, but measurement values are not explicitly corrected for its inﬂuence. Disk rotation results in the formation of two modes for each mode order, one forward and one backward traveling wave. While disk modes theoretically travel at the disk rotational speed, cavity modes travel at the ﬂuid’s rotational speed in the cavity. Assuming bulk ﬂow in the cavity at approximately half the disk’s rotational speed, frequencies of acoustic and structure dominant modes approach each other subject to a change in disk rotational speed. These effects are shown in the following, ﬁrst for structure, then for acoustic dominant modes, and last for speciﬁc mode pairs. For all measurement results of the series “ROTATION”, normalized frequencies f are displayed as deﬁned in Equation (10). In addition to the measurement results, theoretical values for the natural frequencies of structure and acoustic dominant modes at the respective disk rotational speed are given, calculated according to Equations (2) and (6), respectively. Utilized core rotation factors for the acoustic dominant modes are listed in Table 2. As shown in Figure 8, the development of a forward and a backward traveling wave Int. J. Turbomach. Propuls. Power 2022, 7, 15 16 of 22 is clearly visible for each mode, and the measured frequency splits of the disk vibration modes subject to rotation agree well with the theoretical values. Figure 8. Normalized natural frequencies of structure dominant modes in air at 0.1MPa for “slim Figure 8. Normalized natural frequencies of structure dominant modes in air at 0.1MPa for “slim cavity” over circumferential Reynolds number. Experimental and theoretical results. cavity” over circumferential Reynolds number. Experimental and theoretical results. Natural frequencies of the acoustic dominant modes are much less apparent in the Natural frequencies of the acoustic dominant modes are much less apparent in the measur measuremen ement t results results than than the the d disk isk vibr vibration ation modes, modes, as as they they are o are often ften over overlapping lapping wit with h neighboring modes or deflected by them. However, the formation of a forward and a backward traveling wave is evident for the acoustic dominant modes with three, four, and five nodal diameters, as shown in Figure 9. Measured and theoretic frequency splits of the acoustic dominant modes subject to disk rotation do not agree well. Moreover, there is no consistent trend noticeable; measured frequency splits for the (3,0,1) and the (5,0,1) modes are smaller than the theoretical values, while the frequency split of the (4,0,1) mode is slightly bigger. Moreover, core rotation factors adopted to match theoretic values of the three and five nodal diameter acoustic waves with the experimental ones appear to be outside of the usual range. Possibly, the core rotation factors are different from the ones calculated theoretically from Equation (5). Additionally, the number of nodal diameters might influence the natural frequency of the traveling waves, in a way different from that stated in Equations (3) and (6). However, for all analyzed modes, the forward wave ap- pears to exhibit a bigger frequency difference to the non-rotating case than the backward wave. Figure 9. Normalized natural frequencies of acoustic dominant modes in air at 0.1MPa for ”slim cavity” over circumferential Reynolds number. Experimental and theoretical results. Int. J. Turbomach. Propuls. Power 2022, 7, 15 16 of 22 Figure 8. Normalized natural frequencies of structure dominant modes in air at 0.1MPa for “slim cavity” over circumferential Reynolds number. Experimental and theoretical results. Int. J. Turbomach. Propuls. Power 2022, 7, 15 16 of 21 Natural frequencies of the acoustic dominant modes are much less apparent in the measurement results than the disk vibration modes, as they are often overlapping with neighboring modes or deflected by them. However, the formation of a forward and a backward traveling wave is evident for the acoustic dominant modes with three, four, and neighboring modes or deﬂected by them. However, the formation of a forward and a five nodal diameters, as shown in Figure 9. Measured and theoretic frequency splits of the backward traveling wave is evident for the acoustic dominant modes with three, four, and acoustic dominant modes subject to disk rotation do not agree well. Moreover, there is no ﬁve nodal diameters, as shown in Figure 9. Measured and theoretic frequency splits of consistent trend noticeable; measured frequency splits for the (3,0,1) and the (5,0,1) modes the acoustic dominant modes subject to disk rotation do not agree well. Moreover, there are smaller than the theoretical values, while the frequency split of the (4,0,1) mode is is no consistent trend noticeable; measured frequency splits for the (3,0,1) and the (5,0,1) slightly bigger. Moreover, core rotation factors adopted to match theoretic values of the modes are smaller than the theoretical values, while the frequency split of the (4,0,1) mode three and five nodal diameter acoustic waves with the experimental ones appear to be is slightly bigger. Moreover, core rotation factors adopted to match theoretic values of the thr out ee side o andfﬁve the usu nodal al ran diameter ge. Poss acoustic ibly, the waves core rotation fac with the experimental tors are different from ones appear the to ones be outside calculatof ed the theoret usual ica range. lly from Eq Possibly uation , the (5 cor ). Add e rotation itionally, the factors num are dif ber of no ferent fr dal omdiame the ones ters calculated theoretically from Equation (5). Additionally, the number of nodal diameters might influence the natural frequency of the traveling waves, in a way different from that might statedinﬂuence in Equatithe ons (3 natural ) andfr (6 equency ). Howeve of the r, for a traveling ll analyzed waves, mo indes, a way the dif fofer rwar ent d wave from that ap- stated pears in to exh Equations ibit a big (3)gand er fr(6). equen However cy diffe , for rence all analyzed to the non-rot modes, ating the ca forwar se than the b d wavea appears ckward to exhibit a bigger frequency difference to the non-rotating case than the backward wave. wave. Figure 9. Normalized natural frequencies of acoustic dominant modes in air at 0.1MPa for ”slim Figure 9. Normalized natural frequencies of acoustic dominant modes in air at 0.1MPa for ”slim cavity” over circumferential Reynolds number. Experimental and theoretical results. cavity” over circumferential Reynolds number. Experimental and theoretical results. To assess the coupling effect for coupled mode pairs, the respective acoustic and structure dominant modes are presented together in the following. Disk rotation further increases the coupling potential (compared to the standstill values in Table 2), making coupling effects observable. The mode pair with three nodal diameters (displayed in Figure 10) is very well visible for all analyzed rotational speeds. However, the coupling potential is estimated to be low (CP = +0.9), and thus the expected coupling effect relatively small. Theoretically, the forward traveling waves of the structure and acoustic dominant modes would cross at very high disk rotational speeds while the backward waves veer away from each other. However, within the range of analyzed rotational speeds, the coupling potential remains weak. If the theoretic values are adjusted to approximately ﬁt the backward wave of the measured acoustic dominant mode, which is assumed to be less inﬂuenced by coupling, a core rotation factor of about 0.25 must be adopted (as shown in Figure 10). Compared to the backward wave, the acoustic dominant forward wave appears to be deﬂected towards higher frequencies. This might be due to the increased but still relatively small coupling potential at higher rotational speeds. The strongest coupling potential of all measured modes is predicted for the four nodal diameter mode pair in the “wide cavity” test rig setup (CP = +0.08). Therefore, frequency shifts due to the strong coupling effect are expected to be observable. As can be seen in Figure 11, mode veering is visible, conﬁrming this assumption. The forward modes strongly inﬂuence and deﬂect each other. Instead of crossing at a circumferential Reynolds number of approximately 1.7 10 , the forward traveling waves approach each Int. J. Turbomach. Propuls. Power 2022, 7, 15 17 of 22 To assess the coupling effect for coupled mode pairs, the respective acoustic and structure dominant modes are presented together in the following. Disk rotation further increases the coupling potential (compared to the standstill values in Table 2), making coupling effects observable. The mode pair with three nodal diameters (displayed in Fig- ure 10) is very well visible for all analyzed rotational speeds. However, the coupling po- Int. J. Turbomach. Propuls. Power 2022, 7, 15 17 of 21 tential is estimated to be low (𝐶𝑃 = +0.9 ), and thus the expected coupling effect relatively small. Theoretically, the forward traveling waves of the structure and acoustic dominant modes would cross at very high disk rotational speeds while the backward waves veer other away from until a each other. minimumHo normalized wever, within the frequency range o differ f analy ence of zed r 0.2 o is tational spe reached and edsthen , the cou- veer apart, pling p incr oteeasing ntial rem their ainfr s w equency eak. If the separation theoretiagain. c values For are the adju “slim sted cavity” to apptes roxim t rigasetup, tely fit this the behavior backward w cannot ave of th be observed, e measured as aco the acoustic ustic domi dominant nant mod mode e, whis ich shifted is assu towar med to ds be higher less fr infl equencies uenced b ry esulting couplin in g, a a weaker core rot coupling ation facto potential r of abo (u CtP 0.= 25 m +0.28 ust b ). The e adop structur ted (ae s s dominant hown in mode Figure 10). follows Cothe mpared theorto the etical values backwa well, rd wave, the and the acoustic acoustic dominant dominant forw one only ard wave exhibits ap- a slight upward deﬂection. This is a clear indication that a stronger coupling effect is present pears to be deflected towards higher frequencies. This might be due to the increased but for stronger coupled modes, i.e., those in closer proximity to each other. still relatively small coupling potential at higher rotational speeds. Figure 10. Normalized natural frequencies of three nodal diameter mode pair in air at 0.1MPa for Figure 10. Normalized natural frequencies of three nodal diameter mode pair in air at 0.1MPa for Int. J. Turbomach. Propuls. Power 2022“slim , 7, 15 cavity” over circumferential Reynolds number. Experimental and theoretical results (a core 18 of 22 “slim cavity” over circumferential Reynolds number. Experimental and theoretical results (a core rotation factor of 0.25 is adopted for the theoretic values of the acoustic wave). rotation factor of 0.25 is adopted for the theoretic values of the acoustic wave). The strongest coupling potential of all measured modes is predicted for the four nodal diameter mode pair in the “wide cavity” test rig setup ( = +0.08 ). Therefore, frequency shifts due to the strong coupling effect are expected to be observable. As can be seen in Figure 11, mode veering is visible, confirming this assumption. The forward modes strongly influence and deflect each other. Instead of crossing at a circumferential Reynolds number of approximately 1.7 × 10 , the forward traveling waves approach each other until a minimum normalized frequency difference of 0.2 is reached and then veer apart, increasing their frequency separation again. For the “slim cavity” test rig setup, this behavior cannot be observed, as the acoustic dominant mode is shifted towards higher frequencies resulting in a weaker coupling potential ( = +0.28 ). The structure domi- nant mode follows the theoretical values well, and the acoustic dominant one only exhib- its a slight upward deflection. This is a clear indication that a stronger coupling effect is present for stronger coupled modes, i.e., those in closer proximity to each other. Figure 11. Normalized natural frequencies of forward waves of four nodal diameter mode pair in Figure 11. Normalized natural frequencies of forward waves of four nodal diameter mode pair in air air at 0.1 MPa for “wide” and “slim cavity” over circumferential Reynolds number. Experimental at 0.1 MPa for “wide” and “slim cavity” over circumferential Reynolds number. Experimental and and theoretical results (core rotation factors of 0.44 and 0.48 are adopted for the theoretic values of theoretical results (core rotation factors of 0.44 and 0.48 are adopted for the theoretic values of the the acoustic waves for “slim” and “wide cavity” setup, respectively). acoustic waves for “slim” and “wide cavity” setup, respectively). The four and five nodal diameter mode pairs, which are estimated to be strongly coupled in the “slim cavity” setup (𝐶𝑃 values of +0.28 and −0.14, respectively, see Table 2), are not very well visible in the measurement results. With disk rotation, the coupling potential becomes even bigger, but the response peaks of the structure dominant modes are poorly visible and diminish at a certain rotational speed. As shown in Figure 12, the backward wave of the structure dominant five nodal diameter mode disappears once it comes close to the forward acoustic wave and therefore does not come in closer proximity to the potentially coupled backward acoustic dominant wave. The same happens for the forward wave of the structure dominant four nodal diameter mode, which is no longer visible at Reynolds numbers higher than 2 × 10 , where it crosses the backward wave of the acoustic dominant mode (see Figures 8 and 9). Therefore, it does not come close to the coupled forward wave of the acoustic dominant mode. Consequently, no conclusion on the coupling effect and possible veering behavior of the modes with a strong coupling potential can be drawn for the “slim cavity” test rig setup. 𝐶𝑃 𝐶𝑃 Int. J. Turbomach. Propuls. Power 2022, 7, 15 18 of 21 The four and ﬁve nodal diameter mode pairs, which are estimated to be strongly coupled in the “slim cavity” setup (CP values of +0.28 and 0.14, respectively, see Table 2), are not very well visible in the measurement results. With disk rotation, the coupling potential becomes even bigger, but the response peaks of the structure dominant modes are poorly visible and diminish at a certain rotational speed. As shown in Figure 12, the backward wave of the structure dominant ﬁve nodal diameter mode disappears once it comes close to the forward acoustic wave and therefore does not come in closer proximity to the potentially coupled backward acoustic dominant wave. The same happens for the forward wave of the structure dominant four nodal diameter mode, which is no longer visible at Reynolds numbers higher than 2 10 , where it crosses the backward wave of the acoustic dominant mode (see Figures 8 and 9). Therefore, it does not come close to the coupled forward wave of the acoustic dominant mode. Consequently, no conclusion Int. J. Turbomach. Propuls. Power 2022, 7, 15 19 of 22 on the coupling effect and possible veering behavior of the modes with a strong coupling potential can be drawn for the “slim cavity” test rig setup. Figure 12. Normalized natural frequencies of five nodal diameter mode pair in air at 0.1 MPa for Figure 12. Normalized natural frequencies of ﬁve nodal diameter mode pair in air at 0.1 MPa for “slim cavity” over circumferential Reynolds number. Experimental and theoretical results (a core “slim cavity” over circumferential Reynolds number. Experimental and theoretical results (a core rotation factor of 0.30 is adopted for the theoretic values of the acoustic wave). rotation factor of 0.30 is adopted for the theoretic values of the acoustic wave). Nevertheless, Nevertheless, it sho it should uld be be noted noted that that th the e co concept ncept of of bulk bulk ﬂow flow in inside side the the c cavity avity th that at can can be estim be estimated ated by by a a core core rrotation facto otation factorr is is a a theo theorretical etical idea idea th that at does does not not exist exist in re in real al cavities. cavities. Mor Mor eeover over, measurement re , measurement rsults pre esults prsesented ented in th in is wor this work k are ch arear characterized acterized by c by ir- cir cumferen cumferential tial Reyno Reynolds lds numbers slightly numbers slightly lower lower th than an val values ues for for we well-established ll-established bulk bulk fﬂow low. . Therefore, most measurement results actually classify into the transitions area between Therefore, most measurement results actually classify into the transitions area between merged and separated boundary layer regimes according to [25], and the applicability of merged and separated boundary layer regimes according to [25], and the applicability of the core rotation concept should be further investigated. Using a ﬂuid with lower kinematic the core rotation concept should be further investigated. Using a fluid with lower kine- viscosity or conducting measurements with disk rotation at higher pressures could be matic viscosity or conducting measurements with disk rotation at higher pressures could options to reach higher Reynolds numbers in future experiments. be options to reach higher Reynolds numbers in future experiments. 6. Conclusions and Future Work 6. Conclusions and Future Work In the ﬁrst part, fundamental concepts necessary to understand the formation of struc- In the first part, fundamental concepts necessary to understand the formation of tural and acoustic modes and their coupling in a rotor-cavity system are compiled, and structural and acoustic modes and their coupling in a rotor-cavity system are compiled, existing coupling models are reviewed. Based on a comprehensive literature review, an and existing coupling models are reviewed. Based on a comprehensive literature review, overview of frequency shift effects in rotor-cavity systems with a rotating disk is estab- an overview of frequency shift effects in rotor-cavity systems with a rotating disk is estab- lished, including effects for uncoupled and coupled modes. A procedure to analyze those lished, including effects for uncoupled and coupled modes. A procedure to analyze those frequency shift effects is developed and later applied to the analysis of experimental results. frequency shift effects is developed and later applied to the analysis of experimental re- Moreover, an attempt to quantify the expected coupling effects is made by establishing a sults. Moreover, an attempt to quantify the expected coupling effects is made by estab- coupling strength and a coupling potential parameter. lishing a coupling strength and a coupling potential parameter. In the second part of this work, measurement results of acoustic and structure dom- inant modes are presented, including a test series with varying cavity pressure and one with different disk rotational speeds. For the latter, the development of a forward and a backward traveling wave is demonstrated for each mode. All excited mode pairs in the test rig are at least weakly coupled in air, i.e., frequency shifts are observed that cannot solely be explained by “uncoupled mode effects”, such as the added mass, speed of sound, and stiffening effect, but indicate an additional coupling effect. However, the hypothesis of a bigger coupling effect (bigger frequency shift) for stronger coupled modes cannot be corroborated consistently. By varying the disk rotational speed, the frequency difference between coupled modes, and thus their coupling potential, can be varied additionally. However, the expected coupling effect is not observable for all mode pairs. Only the four nodal diameter mode pair in the “wide cavity” exhibits a strong coupling effect. Mode veering is visible, i.e., with an increase in disk rotational speed, the coupled modes ap- proach each other, and then veer apart instead of crossing. Int. J. Turbomach. Propuls. Power 2022, 7, 15 19 of 21 In the second part of this work, measurement results of acoustic and structure dom- inant modes are presented, including a test series with varying cavity pressure and one with different disk rotational speeds. For the latter, the development of a forward and a backward traveling wave is demonstrated for each mode. All excited mode pairs in the test rig are at least weakly coupled in air, i.e., frequency shifts are observed that cannot solely be explained by “uncoupled mode effects”, such as the added mass, speed of sound, and stiffening effect, but indicate an additional coupling effect. However, the hypothesis of a bigger coupling effect (bigger frequency shift) for stronger coupled modes cannot be corrob- orated consistently. By varying the disk rotational speed, the frequency difference between coupled modes, and thus their coupling potential, can be varied additionally. However, the expected coupling effect is not observable for all mode pairs. Only the four nodal diameter mode pair in the “wide cavity” exhibits a strong coupling effect. Mode veering is visible, i.e., with an increase in disk rotational speed, the coupled modes approach each other, and then veer apart instead of crossing. Generally, measurement results for structure dominant modes agree well with theo- retical predictions. However, measured acoustic dominant modes do not always exhibit the expected theoretic behavior. It requires further research to clarify whether or not the underlying assumptions on core rotation apply to the ﬂow conditions in the cavity during the measurements. For future research, a different ﬂuid or a higher pressure could be considered to ensure higher Reynolds numbers and further validate the bulk ﬂow concept as well as examine boundary layer inﬂuences. Additionally, the inﬂuence of the small opening of the cavity needs further investigation. The next essential step in the line of this research will be to investigate the damping of coupled acoustic and structure dominant modes, as resonances cannot always be avoided and the damping ratio determines the response amplitudes of the system and thus the risk of impeller material failure. Eventually, the transferability of the results to real machines needs to be proven. However, this work poses an essential step towards a better under- standing of acoustoelastic coupling effects in the side cavities of centrifugal compressors under varying operating conditions. Author Contributions: Conceptualization, T.U.; Data curation, T.U.; Formal analysis, T.U.; Funding acquisition, D.B.; Investigation, T.U.; Methodology, T.U.; Supervision, D.B.; Visualization, T.U.; Writing—original draft, T.U.; Writing—review & editing, D.B. All authors have read and agreed to the published version of the manuscript. Funding: This project is funded by Deutsche Forschungsgemeinschaft (DFG). The authors are grateful for the ﬁnancial support under project number 429153912. Data Availability Statement: Not applicable. Conﬂicts of Interest: The authors declare no conﬂict of interest. The funders had no role in the design of the study; in the collection, analyses, or interpretation of data; in the writing of the manuscript, or in the decision to publish the results. Nomenclature b disk/impeller thickness, m c speed of sound, m/s c friction coefﬁcient CP coupling potential parameter CS coupling strength parameter f frequency, 1/s f = normalized frequency relative to the frequency f = f (air, 0.1 MPa, 295 K) st,2,0 re f re f f f D f = relative frequency change relative to the frequency f = f 0.1 MPa ( ) 0 Int. J. Turbomach. Propuls. Power 2022, 7, 15 20 of 21 f rotational frequency of rotating disk, 1/s G = non-dimensional axial gap width of cavity k core rotation factor L characteristic length, m m mass, kg (m, n) number of nodal diameters, nodal circles of structure (dominant) mode (p, q, r) number of nodal diameters, nodal circles, axial nodes of acoustic (dominant) mode r radius, m Re circumferential Reynolds number s axial gap width of cavity, m t cylindrical portion of impeller side room, m ax u disk/impeller circumferential velocity, m/s v velocity, m/s W rotational speed of disk, 1/min (pa) dimensionless solutions to the ﬁrst derivative of the Bessel function for the (p,q) p,q acoustic mode n kinematic viscosity of the ﬂuid, m /s r density, kg/m Subscripts ac acoustic (dominant) c, c f , cr cavity, front cavity, rear cavity corr corrected d disk e excitation f ﬂuid s system st structure (dominant) tot total uc uncoupled References 1. Lamb, H.; Southwell, R.V. The vibrations of a spinning disk. Proc. R. Soc. A 1921, 99, 272–280. 2. Southwell, R.V. On the free transverse vibrations of a uniform circular disc clamped at its centre; and on the effects of rotation. Proc. R. Soc. A 1922, 101, 133–153. 3. Campbell, W. Protection of Steam Turbine Disk Wheels from Axial Vibration; General Electric Company: New York, NY, USA, 1924. 4. Eversman, W.; Dodson, R.O., Jr. Free vibration of a centrally clamped spinning circular disk. AIAA J. 1969, 7, 2010–2012. [CrossRef] 5. Ewins, D.J. Vibration characteristics of bladed disc assemblies. J. Mech. Eng. Sci. 1973, 15, 165–186. [CrossRef] 6. Irretier, H. The natural and forced vibrations of a wheel disc. J. Sound Vib. 1983, 87, 161–177. [CrossRef] 7. Irretier, H. Experiments and calculations on the vibrations of rotating radial impellers. J. Vib. Acoust. Stress Reliab. 1988, 110, 137–142. [CrossRef] 8. Tyler, J.M.; Sofrin, T.G. Axial Flow Compressor Noise Studies; No. 620532. SAE Trans. 1962, 70, 309. 9. Ehrich, F.F. Acoustic resonances and multiple pure tone noise in turbomachinery inlets. J. Eng. Power 1969, 91, 253–262. [CrossRef] 10. Eisinger, F.L. Acoustic fatigue of impellers of rotating machinery. J. Press. Vessel. Technol. 2002, 124, 154–160. [CrossRef] 11. Eisinger, F.L.; Sullivan, R.E. Acoustically-induced structural fatigue of impeller discs: A Brief Note. Turbo Expo Power Land Sea Air 2002, 3610, 1063–1069. 12. Magara, Y.; Mitsuhiro, N.; Kazuyuki, Y.; Naohiko, T.; Tetsuya, K. Natural frequencies of centrifugal compressor impellers for high density gas applications. ASME Int. Mech. Eng. Congr. Expo. 2008, 48661, 107–114. 13. Magara, Y.; Yamaguchi, K.; Miura, H.; Takahashi, N.; Narita, M. Natural frequency shift in a centrifugal compressor impeller for high-density gas applications. J. Turbomach. 2013, 135, 011014. [CrossRef] 14. Beirow, B.; Kühhorn, A.; Figaschewsky, F.; Nipkau, J. Effect of mistuning and damping on the forced response of a compressor blisk rotor. ASME Turbo Expo 2015: Turbine Technical Conference and Exposition. Am. Soc. Mech. Eng. Digit. Collect. 2015, 56772, V07BT32A001. 15. Heinrich, C.R.; Kühhorn, A.; Steff, K.; Petry, N. Generalized Model for the Approximation of Coupled Acousto-Mechanical Natural Frequencies in High-Pressure Centrifugal Compressors. J. Eng. Gas Turbines Power 2021, 143, 071022. [CrossRef] 16. Heinrich, C.R.; Unglaube, T.; Beirow, B.; Brillert, D.; Steff, K.; Petry, N. Surrogate Models for The Prediction of Damping Ratios in Coupled Acoustoelastic Rotor-Cavity Systems. In Proceedings of the ASME Turbo Expo 2021: Turbine Technical Conference and Exposition, online, 7–11 June 2021. GT2021-58835. Int. J. Turbomach. Propuls. Power 2022, 7, 15 21 of 21 17. Hellmich, B.; Seume, J.R. Causes of acoustic resonance in a high-speed axial compressor. ASME Turbo Expo 2006: Power for Land, Sea, and Air. Am. Soc. Mech. Eng. Digit. Collect. 2006, 4241, 377–387. 18. König, S. 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International Journal of Turbomachinery, Propulsion and Power – Multidisciplinary Digital Publishing Institute

**Published: ** May 6, 2022

**Keywords: **rotor-cavity system; acoustoelastic modes; coupling; fluid-structure interaction

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