Structural Safety Assessment Based on Stress-Life Fatigue Analysis for T/C Nozzle Ring Blade

Abstract

The performance of the turbocharger nozzle ring is a key factor in the overall operation of the main engine of the ship. Minimizing failure and damage caused by high exhaust gas temperature and pressure is essential. As a first step toward improving turbocharger safety, this study performed 3D scanning of an aged nozzle ring to obtain its precise geometry and developed a corresponding numerical model. The boundary conditions of the numerical model were defined by the exhaust gas temperature and pressure at various engine output loads. Structural safety was assessed using static structural and stress-life fatigue analyses. A sharp increase in maximum equivalent stress and strain was observed at output loads of 85% and higher. At 25% load, the maximum fatigue life indicated $1.76{ \times 10}^8$ cycles, while at 100% load, the maximum damage index reached 1. A field performance test conducted at 85% of the main engine’s output load revealed severe damage under high-load conditions. Specifically, damage occurred at the contact area between the outer hoop and the tip of the blade’s trailing edge. This observed damage pattern closely aligned with the results predicted by the fatigue life analysis. The validity of the present study was confirmed through a comparative analysis of the fatigue life predictions and the field test results.

1. Introduction

In the current landscape of the global economy, over 80% of international trade volume is transported by sea on merchant vessels [1]. To ensure both the speed efficiency of maritime goods transport and the fuel efficiency of marine propulsion, two-stroke diesel engines are predominantly employed as the main engines for ship propulsion. These engines are regarded as one of the most economical options due to their high power output, superior performance, and mechanical efficiency. However, they are also considered a primary contributor to the rise in ship exhaust gas emissions. Consequently, the International Maritime Organization (IMO) has established regulations aimed at controlling emissions to prevent air pollution from ship exhaust and to mitigate climate change. The IMO set limits on nitrogen oxides (NOx), a major air pollutant from ship exhaust, through the implementation of its Tier-III emission standards in 2018. Regarding sulfur oxides (SOx), the IMO mandated a maximum sulfur content in any fuel oil used on board ships and expanded Emission Control Areas (ECA) to more regions, thereby enforcing stricter emission standards [2]. As a result of these rigorous environmental regulations, exhaust gas emissions from ships have been reduced to less than 3% of global greenhouse gas emissions, positioning maritime transport as the cleanest and greenest mode of transportation in comparison to other methods [3].

However, due to maritime accidents such as ship collision, loss of life, engine damage, and marine environmental pollution, the shipping industry is classified as one of the most hazardous sectors [4,5]. An analysis of marine accident events revealed that 25.9% of accidents reported in Europe over the past decade were attributed to system or equipment failures [6]. According to a marine accident statistics report from South Korea, 29.7% of marine accidents were caused by failures in main engines and auxiliary equipment. Given the increasing proportion of marine accidents caused by ship engine damage as reported [7], the Ships Safety Act and Ship Engine Standards of South Korea have stipulated the need for the additional installation of auxiliary equipment. However, there are no mandatory requirements for the additional installation of components related to the main engine, which is central to ship operation. Considering the potential for marine accidents stemming from main engine failure to escalate into secondary accidents such as fires, collisions, and shipwrecks, there has been a growing focus on research concerning the inspection and maintenance of main engines [8].

In particular, turbochargers, an essential component fitted to the main engine, offer high performance efficiency. However, due to the characteristics of high-speed rotating turbines, overspeed operation may occur, leading to reports of numerous accidents causing injury to crew members [5]. Turbocharger failure events pose significant obstacles to the normal operation of ships, and recent studies have been conducted to minimize such accidents and to enhance performance efficiency.

Previous research on turbochargers includes the classification of failure cases based on their causes and the investigation of combustion characteristics [9], as well as analysis of the relationship between main engine design and turbocharger explosions [5]. Additionally, to prevent failure and damage, one study predicted issues caused by dirty air filters or inlet grids using an engine simulator [4], and another focused on diagnosing turbine rotor bearing conditions through analysis of vibration and acoustic signals from the turbocharger [10].

To enhance the performance of turbochargers, nozzle blades have been designed and numerical analyses have been conducted under various flow conditions [11]. Additionally, numerical results of turbochargers have been compared and analyzed against model test outcomes [12]. Studies have also investigated the operational characteristics of turbochargers according to different geometric configurations [13], and optimization of the geometry has been performed [14]. Beyond geometric optimization, various simulation analyses have been carried out to optimize the overall turbocharger system [15,16,17]. In addition to numerical approaches, experimental studies have analyzed the corrosion characteristics caused by bending and torsional stresses at the welded joints of nozzle blades [18,19]. However, most experimental efforts have been limited to analyzing the power characteristics and exhaust emissions of turbocharger systems [20,21,22]. Notably, numerical and experimental investigations into damage of critical rotating components due to high-temperature exhaust gases remain scarce.

Therefore, in this study, aiming to minimize turbocharger failure and damage, a structural safety assessment of the nozzle ring—a component significantly affected by high temperature and high pressure of exhaust gas—was conducted based on static structural analysis and stress-life fatigue analysis. Furthermore, the validity of the numerical analysis results was verified through a field performance test under the actual operating conditions of the main engine. To perform the numerical analysis of the turbocharger nozzle ring, reverse engineering was carried out using a 3D scanner. The surface pressure coefficient and exhaust gas temperature were set as boundary conditions for the finite element model under each engine operating condition. The structural safety of the nozzle ring was evaluated by comprehensively analyzing the fatigue life and damage indices obtained from the stress-life fatigue analysis.

The field performance test of an aged nozzle ring was carried out on a 4700-ton ship built in 2003. This ship was equipped with an aged main engine that had been in operation for more than 20 years. The performance test was conducted in accordance with the operating conditions of the aged main engine, and the results from the numerical model were compared with the outcomes of the field performance test for evaluation.

2. Materials and Methods

The diesel ship used in this study has an overall length of 102.7 m, an overall breadth of 15.6 m, and a depth of 7.3 m, operating at a maximum speed of 17.7 knots. The main engine used in this study is the 6S35MC Mk7 model, manufactured by HYUNDAI B&W (Ulsan, Republic of Korea), as illustrated in Figure 1. This engine features six cylinders with a bore of 350 mm and a stroke of 1400 mm, delivering a maximum continuous rating (MCR) of 6060 bhp, as shown in

Table 1. Specifications of the main engine.

Description	Value
Model	HYUNDAI B&W 6S35MC Mk7
Number of cylinders	6
Cylinder bore	350 mm
Stroke	1400 mm
Output	6060 bhp
Revolution	173 rpm

. The ship’s main engine is equipped with a single vane-flow test rig (VTR) type turbocharger from ABB Co., Ltd. (Singapore, Singapore), a manufacturer renowned for its high reliability in terms of efficiency and maintenance for large diesel engines, as detailed in

Table 2. Specifications of the main engine and turbocharger.

Description	Value
Model	VTR454
Max. rpm	18,420
Max. temperature	590 °C
Weight	3400 kg

. According to the design specifications, the turbocharger has a maximum rotational speed of 18,420 rpm and can withstand exhaust gas temperatures of up to 590 °C.

2.1. Characteristics of Turbocharger

The power output of the main engine is generated through the combustion of ship fuel, and the amount of combustible fuel is determined by the density of air that can be supplied to the combustion chamber. A turbocharger functions as an auxiliary system of the main engine to increase the air density used for combustion, thereby enhancing the power output of the main engine, and torque [18]. Turbochargers are installed in the main engine of ships or in engines used as generators, and multiple turbochargers may be installed depending on the power output requirements of a large main engine [4].

Figure 2 presents a schematic diagram of the VTR type turbocharger used in this study. After fuel combustion occurs in the engine cylinders, the resulting exhaust gas flows from the exhaust manifold at a constant pressure to drive the turbine blades. The compressor on the intake (blow) side, which is mechanically linked to the turbine, compresses the intake air and directs it to the combustion chamber.

In a turbocharging system, depending on the power output and operating characteristics of the engine, either a constant pressure-type or dynamic pressure-type turbocharger is used. The ship examined in this study was equipped with a constant pressure-type turbocharger featuring a large manifold designed to deliver exhaust gas to the turbine at a steady pressure. The constant pressure-type turbocharger is well-suited for high power output engines and offers the advantage of high efficiency by ensuring a stable inflow of exhaust gas. However, its efficiency diminishes at low engine speeds, necessitating the use of an auxiliary blower.

2.2. Design and Characteristics of Nozzle Ring

To maximize efficiency by accelerating the flow of exhaust gas and ambient air directed to the turbine side and blow side of the turbocharger, respectively, a nozzle ring and a diffuser are installed. The nozzle ring functions as a guide blade, converting pressure energy into kinetic energy by accelerating the exhaust gas flow to its maximum velocity before it reaches the turbine blade. Conversely, the diffuser on the blow side is responsible for converting kinetic energy into pressure energy, thereby reducing the friction loss caused by the high velocity of compressed air passing through the impeller.

Figure 3 illustrates the configuration of the nozzle ring along with the cross-section of the nozzle ring blade, showing that the turbine blade has an airfoil geometry. The airfoil-shaped turbine blade results in a change in pressure distribution based on the cross-sectional areas of the inlet and outlet.

Table 3. Mechanical properties of the nozzle ring.

Property	Value
Specific heat	0.5 (J/g°C)
Specific gravity	7.93
Coefficient of thermal expansion	17.3 W/m∙°C
Thermal conductivity	16.3 W/m∙°C
Yield strength	≥175 N/mm2
Tensile strength	≥480 N/mm2
Elongation	≥40%

summarizes the mechanical property values required for the numerical analysis of the nozzle ring used in this study.

2.3. Reverse Engineering for Nozzle Ring

To develop a numerical model based on accurate design information of the nozzle ring, 3D scanning was conducted as part of a reverse engineering process. The performance of the nozzle ring, which converts the pressure energy of exhaust gas from the main engine into kinetic energy, is highly dependent on the structural characteristics of the turbine blades. This is because the exhaust gas flow rate passing through the nozzle ring is determined by the number and spacing of the turbine blades. Specifically, since the turbine blades have an airfoil geometry, the lift-to-drag ratio and performance are influenced by the geometry type and specifications.

For reverse engineering of the nozzle ring, which requires high scan quality for precise capture of design information, the MetraSCAN 3D model from CREAFORM was used. The detailed specifications of the 3D scanner are provided in

Table 4. Specifications of the 3D scanner used for experimental equipment.

Description	Value
Accuracy	0.025 mm
Volumetric accuracy	16.6 m3 (0.078 mm)
Measurement resolution	0.025 mm
Measurement rate	1,800,000 measurement/s
Light source	15 blue laser crosses
Scanning area	310 × 350 mm
Stand-off distance	300 mm
Depth of field	250 mm

. This equipment offers the advantage of allowing the blue laser to accurately reach the overlapping area between the turbine blade’s leading edge and trailing edge of the nozzle ring, enabling precise and sophisticated reverse engineering implementation. Furthermore, the scanner features a stand-off distance display function, making it possible to derive a high-resolution model.

2.4. Numerical Methods

A flowchart illustrating the overall process of numerical analysis in this study is shown in Figure 4. The computational fluid dynamics (CFD) analysis was conducted based on the finite element model of the nozzle ring developed through 3D scanning. The CFX software from ANSYS 2023 was used for the CFD simulations, and the pressure coefficient of the turbine blade surface derived from the CFD analysis, along with the actual temperature of the exhaust gas entering from the exhaust manifold, were applied as the loading conditions for the static structural analysis. The flow field of the numerical model consists of a total of 8.18 million nodes. As shown in Figure 5, a multi-zone method, combining hexahedral and tetrahedral meshes, was used for the flow field mapping of the nozzle ring. For the boundary conditions, the inlet of the nozzle ring flow field was set to a pressure condition representing the uniform inflow of exhaust gas, and the outlet was set to the condition of averaged static pressure. The k-ε model was used as the fundamental turbulence model, while the shear stress transport (SST) model with the k-ω model applied to the boundary layer region of the wall was selected. The SST turbulence model has been widely applied in numerous studies to calculate the adverse pressure on the surface of nozzle ring blades with airfoil shapes [23,24]. Therefore, it enables accurate determination of the pressure coefficient on the nozzle ring blade surface in this study and is known to exhibit superior reproducibility of experimental results compared to other turbulence models [25,26].

Since the nozzle ring of the main engine is located on the turbine side of the turbocharger, static structural analysis was employed to minimize the influence of inertial loads. The equivalent stress value of the nozzle ring, derived from the static structural analysis, was incorporated into the stress-life fatigue analysis of the nozzle ring. The cyclic loading conditions were set to reflect the operating characteristics of the main engine, where the exhaust pressure and temperature for different outputs remain constant. Equation (1) of the cyclic loading condition represents the stress range, while Equations (2) and (3) for ${\sigma }_m$, the mean of the maximum and minimum stress, and stress amplitude ${\sigma }_a$ were used as key factors in the calculation. Previous studies have specifically noted that the value of ${\sigma }_a$ plays a crucial role in determining the fatigue life in stress-life fatigue analysis [27]. In Equation (2), ${2N}_f$denotes reversals to failure, ${{\sigma }^{\prime }}_f$ denotes the fatigue strength coefficient, and b represents the fatigue strength exponent. The load, calculated according to these equations, is applied through the theory of Palmgren–Miner’s linear damage hypothesis, where repeated cyclic loading significantly influences the determination of fatigue damage and life. Fatigue damage can be assessed by calculating the cumulative total damage, D, as shown in Equation (4). The value of D is computed using $N_{fi}$, which represents the number of stress cycles required to reach the fracture point (or failure) under the i-th stress amplitude, and $n_{fi}$, which is the number of cycles under the i-th stress amplitude.

$$\Delta \sigma ={\sigma }_{max}-{\sigma }_{min}$$

(1)

$${\sigma }_a={\displaystyle \frac{∆\sigma }{2}}={{\sigma }^{\prime }}_f·{\left({2N}_f\right)}^b$$

(2)

$${\sigma }_m={\displaystyle \frac{{\sigma }_{max}+{\sigma }_{min}}{2}}$$

(3)

$$\mathrm{D} =\sum {\displaystyle \frac{n_i}{N_i}}={\displaystyle \frac{n_1}{N_{f1}}}+{\displaystyle \frac{n_2}{N_{f2}}}+\dots +{\displaystyle \frac{n_i}{N_{fi}}}=1$$

(4)

2.5. Field Performance Test

The performance of the main engine deteriorates with increasing sailing time and aging [28]. The performance deterioration of the main engine and turbocharger of the ship used in this study is evident from the values presented in

Table 5. Variation in performance of the aged main engine.

	Load (%)	RPM
		Turbocharger	Main Engine
Sea-trial (2003)	75	15,100	157.2
	85	16,100	163.9
Sailing (2024)	75	14,300	147.0
	85	15,400	155.4

. The main engine, which had been in operation for more than 20 years, shows a substantial difference in its performance indicators compared to those measured during the sea trial in 2003. There is a notable decrease in the rpm of the turbocharger compared to the sea trial under the same load conditions, which causes a corresponding decrease in the rpm of the main engine. However, fuel consumption of the main engine increases in an attempt to reach the set rpm and power output, and the exhaust gas temperature also rises, as shown in Figure 6. This increase in exhaust gas temperature is anticipated to significantly impact the structural safety of the nozzle ring. Specifically, the rpm of the sea-trial engine under a 75% load and the current engine under an 85% load are similar, indicating the aging of the main engine of the ship. Therefore, in this study, to reflect the condition of the aged turbocharger, a field performance test was conducted under sailing load conditions of 75% and 85%, as shown in Figure 6. Continuous sailing was performed to compare the results of the field test with those of the stress-life fatigue analysis.

3. Results and Discussion

3.1. Reverse Engineering

The turbine blades of the nozzle ring feature a complex geometry, which presents challenges in accurately capturing the design. Various methods exist for performing reverse engineering on the target object, but in this study, the geometry data of the nozzle ring was acquired using a 3D scanner. The first step in the reverse engineering of the turbine blade involves scanning using a fixed scanner, as shown in Figure 7a. Following this, an optics tracker, as shown in Figure 7b, is used to capture the overlapping airfoil geometry of the turbine blades and the exhaust gas flow path. The optics tracker attaches reference points to the target object and collects 3D data using a cross-pattern multi-line laser, allowing for free movement of the scanner. Figure 8a presents the real-world model of the nozzle ring fitted on the turbocharger used in this study, while Figure 8b shows the 3D numerical model generated through the reverse engineering process. Reference points are attached to the surface of the turbine blades to utilize the optics tracker functionality, ensuring that the complex geometry of the blades and flow path is accurately reconstructed.

3.2. Numerical Analysis

3.2.1. Load Distribution

Deriving the load distribution through CFD analysis is a crucial step in the process of structural safety assessment, and the load values are applied to the surface of the nozzle ring to proceed with the analysis. The load on the turbine blade surfaces of the nozzle ring is assigned to each local area, taking into account variations in orientation and magnitude. Figure 9 compares the pressure distributions at 75% and 85% output loads, analyzed through both numerical analysis and field performance tests, segmented into inlet and outlet. The maximum surface pressure coefficients derived from the calculations were approximately 0.464 MPa and 0.613 MPa, respectively. Figure 10 illustrates the pressure distribution on the nozzle ring surface and the exhaust gas temperature measured during the main engine operation. At a 100% output load, a surface pressure coefficient of 0.272 MPa was calculated, and the exhaust gas temperature was recorded at 460 °C. In both cases, the flange, inner hoop, and outer hoop fixed to the turbocharger exhibit minimal change in surface pressure coefficients under different output load conditions, and the nozzle ring inlet area demonstrates a uniform pressure distribution, regardless of engine load. The outlet area of the nozzle ring shows a higher pressure coefficient at output load 85% compared to output load 75% overall. Additionally, as the exhaust gas passes through the turbine blades, pressure energy is converted into kinetic energy, and the pressure coefficient at the trailing edge exhibits lower values, which is attributed to the airfoil geometry applied to the turbine blades.

To analyze the sensitivity of the numerical analysis model, a mesh dependency evaluation was conducted. If the results vary significantly with mesh size, the reliability of the numerical model cannot be guaranteed; therefore, an appropriate mesh must be established. In this study, six mesh models with varying mesh sizes were constructed, as summarized in

Table 6. Mesh independence test of the nozzle ring.

Mesh Size (mm)	Nodes	Elements
1	3,002,917	2,453,090
2	1,448,491	883,708
3	974,867	595,281
5	740,385	454,335
7	646,045	395,331
10	611,681	374,871

. As shown in Figure 11, Static structural analyses were performed using these mesh models, and sensitivity was assessed by examining changes in equivalent stress and strain. The models with mesh sizes of 5, 7, and 10 mm exhibited relatively large variations compared to the results of the finer mesh models. In contrast, the models with mesh sizes of 1, 2, and 3 mm showed nearly consistent results. Accordingly, the mesh model with a size of 3 mm, which falls within the range of low sensitivity to mesh size, was selected for this study.

3.2.2. Stress-Life Fatigue Analysis

The structural performance of the nozzle ring used in this study is significantly influenced by repeated exposure to thermal–mechanical stress. As the internal components of the main engine age, they are subjected to higher temperatures and pressures than under normal operation to achieve high efficiency and power output. Exposure to such extreme environments can lead to the deterioration of the material strength of the internal components of the turbocharger, and a proper investigation into the prediction of structural safety is essential to ensure stable operation.

Before conducting fatigue analysis under thermal-mechanical stress, the structural stability of the nozzle ring was assessed using static structural analysis. The parameters considered for the static structural analysis included blade deflection, maximum equivalent stress, and maximum strain. The boundary conditions were defined based on the surface pressure coefficient of the turbine blades at different output loads of the main engine. Figure 12a shows the maximum blade deflection of 0.47 mm at 100% output load. Figure 12b illustrates the distribution of the maximum equivalent stress and maximum strain. All three parameters demonstrated an overall increase as the output load increased. Notably, the equivalent stress and strain exhibited a sharp increase after reaching 75% output load, which is likely attributed to the significant rise in exhaust gas temperature. Furthermore, the blade deflection and strain values remained within the range of mechanical properties, but the maximum equivalent stress at 100% output load exceeded 8000 MPa. The maximum equivalent stress reflects the magnitude of the real stress and was used in the stress-based fatigue analysis.

For an in-depth investigation of the local characteristics of the turbine blades, the airfoil cross-section of the turbine blade was normalized by the chord length of the leading edge, as shown in Figure 13, with c/C = 0.1, 0.5, and 1.0. The areas through which the exhaust gas flows at the inlet and outlet differ due to the airfoil geometry, resulting in varying surface loads on the pressure surface and suction surface. Consequently, the distribution of equivalent stresses on the pressure and suction surfaces of the turbine blade at 85% output load is illustrated in Figure 13b,c. The equivalent stress on both the suction surface and the pressure surface is higher at the edge areas, where the turbine blade contacts the inner and outer hoops, compared to the center areas. Specifically, the maximum equivalent stress occurs in the bottom tip area where the turbine blade meets the outer hoop. Moreover, due to the high-stress distribution in the area in contact with the outer hoop, bending deflection is likely to occur. This leads to predicted deflection and failure due to stress in the trailing edge, where the turbine blade is thinner compared to the leading edge.

Rapid changes in the rpm of the main engine, resulting from frequent berthing and departure of ships, lead to high cycle fatigue conditions in turbine blades. In particular, high cycle fatigue damage of SUS 316 in the turbine blade of the nozzle ring constitutes a significant percentage of marine accidents [29]. Therefore, a stress-based fatigue life analysis was conducted to predict and prevent damage to the turbine blades from high cycle fatigue. The S-N curve, illustrating the number of cycles to failure of the material used in the nozzle ring due to alternating stress, is presented in Figure 14. The alternating stress is the stress induced by repeated cycles applied until material failure, and its magnitude is represented as ${\sigma }_a$ in Equation (3) for the mean stress. In the stress-life fatigue model, fatigue types are classified into high cycle fatigue and low cycle fatigue based on the magnitude of the applied stress and the yield point of the material. In the static structural analysis, the maximum equivalent stress of the nozzle ring blades exceeds the yield point, but this is limited to specific areas where the outer hoop and the blade are in contact. Additionally, since the blade deflection is minimal due to the characteristics of the nozzle ring, as shown in Figure 12a, this corresponds to the high cycle fatigue condition commonly encountered. High cycle fatigue theory aims to predict fatigue life and damage based on long-term, repeated stress within the elastic limit. The S-N curve exhibits a sharp decrease in slope from ${10}^4$ cycles, indicating the endurance limit of the material at ${10}^5$ cycles. Therefore, the stress-life fatigue analysis was based on the high cycle threshold of ${10}^4$ as the cut-off. The fatigue life analysis based on the S-N curve were conducted under fully reversed loading conditions, with a stress ratio (R) of −1 and a mean stress of 0. To account for the influence of mean stress in fatigue life analysis, Goodman’s theory was applied, considering the structural sensitivity of the turbocharger nozzle ring to tensile stress effects.

The results of the stress-life fatigue analysis are evaluated in terms of fatigue life and fatigue damage index. Figure 15a illustrates the variation in the damage index with respect to the number of cycles to failure. Damage occurs at a number of cycles less than or equal to ${10}^{7.2}$, which corresponds to output loads of 85% and 100%. Figure 15b plotting the distribution of fatigue life and damage values according to changes in the output load. The fatigue life curve shows a maximum of $1.76\times {10}^8$ at 25% output load, while the damage index curves reach 1 at 85% output load. The fatigue life curve shows a decreasing trend with increasing output load, while the damage index shows the opposite trend. However, for both the fatigue life and damage indices, the slope in the curves increases concurrently from output loads of 75% and above. This trend is believed to result from increased fuel consumption of the main engine at output loads of 85% and above, leading to an increase in exhaust gas temperature and surface pressure coefficient.

The fatigue life and damage of the nozzle ring blade with changes in output load can be observed in more detail in Figure 16 and Figure 17. The middle area between the leading edge and trailing edge at 75% and 85% output load shows relatively high fatigue life. However, at 85% output load, the fatigue life in the edge areas of the leading and trailing edges of the blade decreased to nearly ‘0’, which is believed to be caused by the internal stress generated during blade deflection due to repeated stress. Additionally, the damage to the nozzle ring follows a similar trend to the fatigue life, as shown in Figure 16. However, in contrast to the fatigue life, only the edge areas of the blade in contact with the outer hoop show significant damage, with the damage index values at the leading edge and trailing edge remaining small.

3.3. Field Performance Test

The performance indicators for the main engine of the ship show significant differences from those measured during the sea-trial period due to aged equipment and components. Consequently, the main engine is operated with an output load of 75% set as the navigation full load to ensure stable operation. Therefore, a field performance test with the main engine output load of 75% was performed to verify the reproducibility of the results from the stress-life fatigue analysis, as described in the previous sections. Additionally, since most of the performance indicators from the numerical analysis exhibited sudden changes at an output load of 85%, a field performance test was conducted to include the 85% output load case alongside the 75% output load case. Figure 18 shows the performance indicators of the main engine measured during the field performance test, illustrating the concurrent increase in fuel consumption required to follow the commanded rpm and power output. The scavenge pressure of the main engine and the inlet temperature of the turbocharger measured at output loads of 75% and 85% increased by more than 0.5 kg/cm² and 20 °C, respectively, while the measured cylinder exhaust gas temperature reached 400 °C.

Before performing the field performance test, an opening inspection was conducted as shown in Figure 19 to check the condition of the turbocharger nozzle ring. Image (a) shows the compressor wheel mounted on the blow side of the turbocharger, and (b) shows the turbine wheel and nozzle ring on the turbine side. The turbocharger, which was operated stably at a normal output load of 75%, had good components, including the compressor wheel, turbine wheel, and nozzle ring. In particular, no cracks were found in the blades located in the exhaust gas outlet and inlet areas of the nozzle ring.

However, the field performance test at 85% output load showed results that were in contrast to those of the 75% output load. During the 85% output load test, surging and loud noise occurred on the turbocharger side. Following an opening inspection of the turbocharger to determine the cause, a damaged nozzle ring was found, as shown in Figure 20. The nozzle ring of the turbocharger, which had been continuously exposed to the 85% output load for more than 60 h, was seriously damaged. Vibration and noise persisted even after the output load was restored to its original state and the turbocharger was operated again. Figure 20 shows the actual appearance of the damaged nozzle ring, and a 3D model of the damaged nozzle ring was created using a scanner to perform a detailed cause analysis.

The real nozzle ring blade showed severe damage in the area in contact with the surface of the outer hoop, which closely resembled the area where the maximum equivalent stress and maximum damage index were derived from the stress-life fatigue analysis. In particular, the tip region of the blade trailing edge in Figure 20 exhibited more extensive damage due to thermal—mechanical stress. Fatigue cracking that occurred in the area where the outer hoop and the tip of the blade trailing edge were in contact propagated toward the leading edge, further advancing the crack. The propagated crack reduced the cross-section of the blade tip and extended toward the inner hoop side, although it is believed that fracture of the entire blade did not occur. In contrast, no cracks were detected on the surface of the inner hoop area, which remained intact compared to the outer hoop.

Additionally, cracks did not develop in the leading edge of the nozzle ring blade where the exhaust gas from the main engine flows through. However, the outer hoop area of the trailing edge showed serious damage and numerous sites of pitting on the surface. It is believed that the significant damage in the trailing edge area, compared to the leading edge, resulted from the large deflection of the blade and the characteristics of the nozzle ring blades formed with thin thickness. It is reasoned that the repeated thermal–mechanical stress had a dominant effect on the structural performance of the nozzle ring blades.

4. Conclusions

In this study, reverse engineering was performed using a 3D scanner for the analysis and assessment of the structural safety of an aged turbocharger nozzle ring. A numerical model was developed, and static structural analysis and stress-life fatigue analysis were performed under real-world operating conditions. To test and verify the validity of the numerical analysis, a field performance test was conducted, and the test results were compared with the analysis. The main findings derived from the study are summarized as follows:

(1)

A 3D scanner was used to develop a numerical model for the nozzle ring of the turbocharger with complex geometry. Since the performance of the nozzle ring is determined by the geometric characteristics of the blade and flow path, an optics tracker was used to acquire geometric data, which formed the basis for accurate analysis in the numerical study. A numerical model was derived accordingly.

(2)

For the analysis and assessment of the structural safety of the nozzle ring reproduced through reverse engineering, the surface pressure coefficient and exhaust gas temperature for each output load were set as boundary conditions in the finite element model of the nozzle ring. The numerical modeling confirmed exposure to higher exhaust gas temperatures and pressures for the aged main engine to achieve its original performance. Static structural analysis, with the surface pressure coefficient and exhaust gas temperature set as boundary conditions, resulted in a maximum blade deflection of 0.47 mm, and a sharp increase in the maximum equivalent stress and strain was observed at output loads of 85% or above.

(3)

The stress-life fatigue analysis of the nozzle ring was performed based on the S-N curves and the high cycle fatigue theory, considering the operating characteristics of the ship, with a ${10}^4$ cycle set as the cut-off in the analysis. The fatigue life derived from the stress-life fatigue analysis showed a maximum of $1.76\times {10}^8$ cycles at 25% output load, while the damage index was $1.40\times {10}^{31}$ at 100% output load, showing conflicting trends.

(4)

The fatigue life and damage of the nozzle ring blade exhibited locally different characteristics between the leading edge and trailing edge. At 85% output load, the fatigue life was high in the middle area between the leading and trailing edges, but at the edge area near the outer hoop, the fatigue life was close to ‘0’. This is reasoned to be attributable to the concentration of damage in the area, as the high internal stress of the blade was repeatedly applied, as derived from the static structural analysis. Additionally, the damage index of the nozzle ring was small throughout the leading edge and trailing edge areas, except for the edge area of the blade.

(5)

A field performance test of the main engine was conducted to evaluate the validity of the numerical analysis results under the operating conditions of output load 75% and output load 85%, the latter corresponding to an indicated cylinder exhaust gas temperature exceeding 400 °C. At 85% output load, severe damage occurred in the nozzle ring blade, and propagation of the fatigue crack in the area where the outer hoop and the tip of the blade trailing edge were in contact, extending toward the leading edge, was confirmed. The area of damage observed in the field performance test and the area of maximum damage index from the stress-life fatigue analysis were highly similar, verifying the validity of the numerical analysis in this study.

This study investigated the structural integrity of a turbocharger nozzle ring through fatigue analysis. To prevent failures of aged nozzle rings in the future, it is considered necessary to conduct further research aimed at establishing an integrated predictive maintenance system by analyzing various data.