1. Introduction

2. Methods

2.2. Carbon Composite Blades

The final blade and overall engine geometry of the $16th$ iteration were afterwards transferred into a manufacturable design maintaining the aerodynamic surfaces and geometries. Apart from the composite blades, discussed in the following, the design of further components is summarized later. The most significant advantage of using composite structures lies in the substantial reduction of weight, up to $70-80\%$, compared to traditional metallic blades. This reduction enables highly dynamic behavior and enhances the engine’s gravimetric power density. Additionally, there is an increase in the blade’s safety factor and a potential decrease of up to $40\%$ in the maximum relative profile thickness, leading to aerodynamic advantages [30,31]. However, despite these advantages, the widespread adoption of composite blades faces certain challenges. Issues related to repairability and recycling remain open questions. Furthermore, the manufacturing process for composite blades is more complex and labor-intensive compared to traditional milling methods, relying heavily on manual craftsmanship. Consequently, the application of composite blades is currently limited and predominantly reserved for a few large-scale projects.

While research projects of a similar scale to the ELAPSED project are scarce, one noteworthy example is the counter-rotating integrated shrouded propfan (CRISP) project conducted by DLR [32]. This project utilizes a standard structure of the individual composite plies for the two rotors, which deliver a pressure ratio of $1.3$ at mass flow rate of $130kg/s$. This employed structure adheres to a typical symmetrical layup of composite plies, allowing for quasi-isotropic properties. It consists of a symmetrical arrangement of 128 plies, following the $0^{\circ }/\pm {45}^{\circ }/{90}^{\circ }$ orientation, forming an organo sheet. This sheet is subsequently subjected to hot forming and machined to achieve the desired contour. The primary load considerations for these blades, however, are radial and torsional stresses. Therefore, within the CRISP project, various approaches to numerically optimize the layup configuration are being explored, enabling a reduction of the maximum displacement by up to $38\%$ [33].

For the ELAPSED project, however, two approaches have been developed to enable the adoption of this technology in smaller research endeavors. Firstly, the development of a new design tool, as previously explained, has streamlined the direct and automated integration of both composite layers and moulding creation. This represents a significant departure from the previously intricate and time-consuming process of generating these layers and mouldings for existing designs created in different programs. In this context, where the contour or outer geometry of the blade is precisely known in the form of a point cloud and does not need to be separately imported or recreated, the design tool allows for the flexible calculation of the maximum local ply count based on the to be specified ply thickness and maximum relative blade thickness, both being input parameters for the design tool. Additionally, a fundamental layup configuration is embedded within the code. Since a common configuration of individual blades mounted on an inner disk was chosen, the blade consists not only of the actual blade, its upper part, but it includes a blade root as well. This lower part has a front and rear nose that is used to mount the blades into the disk’s lugs.

Exemplarily, the resulting ply structure of the blade’s tip is shown in the left-hand part of Figure 4. It consists not only of the two outer plies, predefined by the aerodynamic surface, but also out of two inner layers on both suction and pressure side, resulting in a total of six layers in its center. Since the blade’s maximum thickness increases towards the hub, the number of plies in the transition from the actual blade to the root accounts to 12. As shown in the center of Figure 4, all layers do not end at the transition to the root but are wrapped around a carbon fleece utilized in the root. The layers are finally unwrapped within the CAD software to receive the individual $2D$ ply shown on the right-hand side of Figure 4. All these steps described so far are included in the design tool described above and are processed automatically.

The second approach developed in this context involves the utilization of cutting-edge stereolithography (SLA) $3D$ printing technology. Due to the intricate blade geometry, conventional composite manufacturing methods such as autoclave production were not feasible options. Instead, the design tool automatically generates five moulding parts. Apart from the two main parts for both pressure and suction side, based on the blade geometry similar to the creation of the individual plies, three further parts are used for the blade’s root. To replace the costly and time-intensive milling procedures in the production of these mouldings, several $3D$ printing options were explored. However, due to the requirement for a smooth surface finish, fused deposition modeling (FDM) printers did not meet the criteria. Instead, SLA printers were employed, reducing the labor involved in moulding production significantly as no further manual post-processing is required. The mouldings for the reference propulsion system blades were still produced using glass reinforced resins. This material meets the strength requirements for moulding pressing but exhibited high brittleness. Consequently, the mouldings could only be used 2-3 times, and occasional chipping from the mouldings resulted in rejected blades. However, through systematic variation of post-curing duration and temperature on newer and even more cost-effective resins, followed by tensile and notch impact tests, the ductility and maximal elongation could be increased significantly by factor five, even offering a smoother surface finish. The initial version with mouldings printed by an FDM printer, resulting in the uneven surface discussed above, compared to the final version of an RPS blade with a smooth surface, is presented in Figure 5. Since both blades are not used within the rotor their small flaws on the surface are not post-treated by filling them with a mixture of resign and ground carbon fiber.

Several tested prepreg materials did not yield satisfactory results because of the small radii within the transition zone. Instead, traditional dry $160g{m}^{-2}$ twill fabric is employed. It is impregnated with the infusion resin MGS RIMR426, cut to match the $2D$ plies discussed earlier, and positioned into the mouldings via hand-laminating. To accommodate torsion, the outermost plies are oriented at $\pm {45}^{\circ }$, while all inner plies are oriented at $\pm {90}^{\circ }$ relative to the radial direction. Subsequently, the five-part moulding is pressed together, and the blade cures initially at room temperature before being relieved out of the mouldings and tempered by ${100}^{\circ }C$ for another hour. Due to the anisotropic material behavior and the consequent need to model the individual plies and their interaction in an FEM analysis, such structural analysis regarding elongation and strength is considerably more complex compared to metallic materials, for example. Consequently, for the RPS, which experiences relatively low loads, this analysis was omitted and the already discussed ply orientation was chosen based on the findings in [33]. Instead, the tensile tests discussed below are intended to be used in the future for validating such a detailed FEM model. This validated model will then be employed to calculate the significantly more heavily loaded blades of the IPS prior to their production. However, different post-curing conditions of the blades were tested, resulting in the optimal tempering discussed above.

Apart from the eleven blades produced for the actual rotor, eighteen further blades were manufactured. Those were used, on the one hand, for initial tests regarding the printer selection, based on the achievable surface smoothness. On the other hand, they were used for the following tensile and bending tests, including a variation of the blade’s curing conditions. Since the overall production time could be reduced to 2 hours per blade, this process seems to be an applicable option within the university’s context. However, the rotor should be used to validate the design tool. Consequently, any geometric deviations from the CAD could prove critical, potentially leading to discrepancies in behavior compared to the CFD calculations. Compared to the target maximum thickness of $1.93mm$ at the blades’ tip, the respective mean value measured accounts to $1.97mm$ having a standard deviation of $0.05mm$. Besides, the chord is measured to $57.63mm$ at the blade’s tip with a standard deviation of $0.83mm$ whereas the target was $59.11mm$. This will be revisited during the tool validation later. To compensate for inequalities in the radial length of the blades the entire rotor with the finally assembled blades was machined down to the desired outer radius in one step, with the radius determined based on the tensile tests and commissioning tests discussed in detail below. In addition to ensuring geometric reproducibility, which could thus be guaranteed, reproducibility of blade mass for balancing is crucial for rotor operation. Therefore, at least a static balancing could be carried out successfully ensuring that the rotor’s center of gravity lies within the engine axis. A balancing mass was applied under the base of all blades scheduled for the rotor operation after curing. This mass consists of the same resin material already used for the actual plies but was reinforced here with ground carbon fiber. Subsequently, material was removed from the balancing mass again using a precision scale having an accuracy of $0.1mg$, resulting in an average blade mass $m_b$ of $17925.9mg$ with a very low standard deviation of $2.5mg$.

As previously mentioned, the tip clearance between the blade and casing at design point speed ${\mathrm{TC}}_{DP}$ is one of the input parameters for the design tool. While a smaller value enhances the blade’s performance by reducing reverse flow and swirl effects, a certain margin must be maintained to prevent grinding of the blades at the casing, for instance during maneuvers resulting in unexpected blade elongation effects. Given the anticipated relatively high E-modulus for composite blades, a moderate ${\mathrm{TC}}_{DP}$ of $0.5mm$ has been selected. However, due to the radial loads caused by rotation, blade tip grinding is necessary to establish a specific ${\mathrm{TC}}_0$ at $n=0rpm$, which then reduces to the chosen ${\mathrm{TC}}_{DP}$ with increasing engine speed. The blade elongation causing the reduction in tip clearance depends mainly on the material’s Young’s modulus E. Since composite belongs to the category of anisotropic materials, its Young’s modulus in a particular considered direction, in this case the radial direction, is heavily influenced by the layer sequence and structure. Unlike isotropic materials, performing an FEM calculation to determine the anticipated elongation is projected to be quite resource intensive. Furthermore, since this approach necessitates subsequent validation owing to the multitude of boundary conditions that need to be set, this step is omitted. Instead, the blade’s performance is experimentally validated both by tensile and operational tests, discussed in the following, and the findings are leveraged to determine an expected blade elongation and a suitable ${\mathrm{TC}}_0$.

As a basis for defining an appropriate testing procedure, a calculation of the expected blade loading is carried out. The blades are exposed to the centrifugal and aerodynamic forces. While the latter load is indicated by the subindex $"a"$, a subindex for indicating centrifugal forces is omitted in the following for the sake of improved readability. First, the centrifugal force F caused by the actual blade is calculated, excluding the root since it is clamped into the disk. At any radial position $r_j$ between hub and shroud, $r_{hub}$ and $r_{shroud}$, respectively, the effective force can be determined by integration as follows:

$$F(\omega ,r_j)=-{\omega }^2·\rho {\int }_{r_{shroud}}^{r_j}A\left(r\right)·rdr.$$

(2)

The blade’s cross section $A\left(r\right)$ varies across the radius and is therefore set by the CAD data. Besides, a constant density $\rho$ across the entire blade is assumed, calculated from the measured mass and the $CAD$ model volume. According to this approach, the maximum centrifugal force occurs at the transition into the foot:

$$\begin{array}{c}\hfill F({\omega }_{DP},r_{hub})=1555N\end{array}$$

(3)

where the angular velocity ${\omega }_{DP}$ corresponds to the design point speed of $14487rpm$. With increasing radius it diminishes, reaching zero at the blade’s tip. Though, in contrast, the aerodynamic axial force of $14.68N$ at design point, determined from the CFD simulation, is much smaller, the blade behavior under this load is experimentally tested, too.

Since the samples are not standardized, the testing procedure for the tensile test must be conducted in accordance with DIN EN ISO 527-5 [34]. To facilitate the experiments in academic research, this test setup was chosen as it is designed for a universal testing machine capable of operating in both tension and compression directions. The model used here for both tests is the inspekt table 50 kN by Hegewald & Peschke. It features wedge-screw grips on both the lower fixed and the upper movable traverse. The lower clamping of the blade is achieved by using a two-part holder, similar to the engine’s disk, which is bolted to the lower traverse. An adapter for the upper traverse is utilized to avoid torsional moments on the blade caused by its twisted geometry. To clamp the curved and twisted tip surface using the jaws of the clamping tool, flat outer surfaces of the elements used to apply the forces are required. To mill the negative profiles of the blade’s pressure and suction side, the manufacturing of these elements as depicted in DIN EN ISO 527-5 is only partially feasible. Instead, their thickness had to be increased and the use of cross-laminated glass-fiber reinforced plastics (GFRP) fabrics is unsuitable due to the cutting of a significant portion of the fibers during the manufacturing process, leading to the loss of required properties. Given that DIN EN ISO 527-5 mandates that the strength values and coefficient of variation must be at least equivalent to those of GFRP, the commonly used aviation-grade aluminum 7075 is employed. The stipulated radial extension of those profiles, $50mm$, is reduced to $20mm$ to cover as little of the blade as possible while avoiding slipping. To enhance the coefficient of friction at the contact interface, the elements are subjected to SiC sandblasting. There is no adhesive bonding between the blades and the described elements. It is worth mentioning that the pressure side element is additionally employed in the bending test to achieve a uniform load distribution on the blade tip.

This approach ensures that all carbon fibers are effectively stressed, thus closely simulating engine operational conditions.

The subsequently discussed test series was conducted on blades of identical design and curing conditions mentioned above and like the ones actually installed on the engine. In Figure 6, the measured absolute elongation $\Delta s_i$ is plotted in dependency of the blade i and the applied force F. Upon examining this diagram, two distinct gradient regions become evident. While the slope is relatively shallow and intermittently wavering for forces $F<2000N$, a linear, steeper increase is observed above this threshold. The lower slope likely results from preload compensation. Reviewing the videos and photos taken during the experiments indicate additional slipping effects of the blade within its clamping, causing sudden, minor decreases of the applied load.

In general, all blades withstood the expected maximum operational load at design point speed, given above in equation 3. In all cases that a damage pattern became visible, causing the final, massive load decrease, it could be classified as delamination at the transition from blade to blade root at the trailing edge (see Figure 7).

Therefore, the safety factors S summarized in the first line of Table 5 are calculated from the maximum measured load, in dependency of the lowest and highest strength blade as well as averaged over all blades, and the estimated operational load at the transition between blade and root, determined above in equation 3.

Noteworthy is the case of blade number 6, considered the most critical scenario due to its notable deviation from the other samples, achieving a maximum of $F_{max}=3280N$ only. Nonetheless, it maintains a safety factor of $2.11$, allowing it to withstand. If the rotational speed is increased to $120\%$, the increased load estimation of $F=2239N$ decreases this safety factor to $1.47$ only, allowing it to withstand, albeit with a marginal reserve. However, a safety factor of $S_{bending}>14.9$ could be determined within the bending test, derived from the most conservative force calculation at the design point.

Returning to the tensile testing, a specific Young’s modulus cannot be calculated since the blade’s cross-section varies across its radial extension whereas the elongation $\Delta s$ was measured over the entire blade only. However, averaged slopes ${\left(\frac{ds}{dFdr}\right)}_i$ can be determined from the measured elongation and load as well as the radial extension of the blade $\Delta r=r_c-r_{hub}$ where $r_c$ represents the radius of the outer clamping. These slopes are separated into two cases as discussed above. On the one hand, the lower gradient ${\left(\frac{ds}{dFdr}\right)}_l$ represents the initial, shallow slope measured during tensile tests. On the other hand, the upper, steeper slope is described by ${\left(\frac{ds}{dFdr}\right)}_u$, based on the assumption that the clamping effects do not occur during operation. Both cases are evaluated, summarized in the lower part of Table 5, and utilized to calculate the expectable blade elongation $\Delta s$ as follows:

$$\begin{array}{c}\hfill \Delta s(\omega ,r_j)={\left(\frac{ds}{dFdr}\right)}_i·{\int }_{r_{hub}}^{r_j}F(\omega ,r)dr\end{array}$$

(4)

assuming that the considered ${\left(\frac{ds}{dFdr}\right)}_i$ is the same for all radial positions and applying equation 2 to calculate the individual elongations of each increment $dr$ at radius r. The resulting tip clearances in dependency of the relative engine speed with regard to the design point speed are plotted in Figure 9, evaluated for an initial ${\mathrm{TC}}_0$ of $1mm$ at installation condition. The error bars are formed by the results of the lowest and highest strength blade, respectively. It can be inferred that at $n_{DP}=100\%$ an elongation between $0.49mm$ and $0.23mm$ is present on average, considering the lower and upper slope.

These results pave the way for engine testing in terms of a tip clearance measurement. The utilized high-speed camera FASTCAM Mini AX200 was positioned in front of the engine. For these tests, the engine was not equipped with its standard inlet but with a similar one having a smaller axial length. This enabled to capture the tip clearance within the red square in dependency of the engine speed as shown in the upper part of Figure 9. A frame rate of $120kfps$, a zoom lens and an additional lighting source behind the engine facilitated analyzable shots. From the videos of the individual speeds those photos were selected where the individual blades are at the same circumferential position to ensure comparability. The actual tip clearance was determined from the calibrated length between two reference positions at the casing and at the blade’s tip, exemplary shown in the bottom part of Figure 8.

Figure 8. Tip clearance quantification by a high-speed camera.

Figure 8. Tip clearance quantification by a high-speed camera.

Figure 9. Experimentally measured tip clearance in dependency of the actual engine speed n, compared to the expected results from the tensile tests based on Table 5.

Figure 9. Experimentally measured tip clearance in dependency of the actual engine speed n, compared to the expected results from the tensile tests based on Table 5.

Besides, the captured videos were evaluated for critical flutter effects, differing elongations of the individual blades or misbalancing not detectible within the former static balancing. However, all these tests were negative.

In order to definitely avoid contact to the casing, a conservative grinding of the blades’ tips was conducted to achieve a tip clearance ${\mathrm{TC}}_0$ at installation condition of $1mm$.

The actual measured tip clearance and a resulting, exponential fitted curve are included in Figure 9 as solid and dotted red line. On the one hand, there is a good agreement between the measured elongation and the calculated values based on the upper, larger gradient measured by the tensile tests though the calculations seem to underestimate the actual elongation slightly. Three possible assumptions could be considered for this deviation. First, it appears reasonable that the the density within the actual blade area is higher than that of the root due to the fleece used within it. Second, the blades are slightly thicker than initially designed as discussed previously. Both would result in higher centrifugal forces and therefore larger elongations when assuming that the deviating thickness results from additional resin which only increases mass but does not bear significant loads. In addition, the calculated slopes are averaged over both the actual blade and the transition zone. Due to the strong curvature of the transition in comparison with the plane structure of the actual blade it can be expected that their slopes vary widely. On the other hand, the resulting tip clearance at design point, approximately $0.74mm$, is larger than the initially specified value of $0.5mm$ and further discussed below within the context of the CFD-validation.

Nevertheless, the test procedure seems to be a suitable indicator to predict the actual operational behavior. Even if the potential range of expectable tip clearances appears to be large and requires careful commissioning tests of the engine including a tip clearance measurement, the calculations could be used to determine more appropriate values for ${\mathrm{TC}}_0$ than set here. To improve the calculation accuracy further, the elongations of the different zones should be measured individually due to their substantial difference in ply structure, for instance by optical measurement systems. Besides, elongations of the inner disk were excluded and may account for deviations between tensile test and experimental results, too. Addressing the second research question raised above, it can be summarized that the design, manufacturing and commissioning of composite-blades became feasable. Key factors are the direct integration of the manufacturing procedure into the design tool itself and the utilization of innovative approaches like the $3D$ printed mouldings, reducing the manufacturing effort massively. Besides, the safety factors calculated above indicate a safe operation over a wide range of engine speeds and the test procedure could be verified. However, the higher total pressure ratio of the future contra-rotating propulsor requires larger engine speeds. To account for the subsequent higher loads, the blade design, especially the layer structure, should be optimized. To improve the design procedure, the test results could be utilized to set up a detailed FEM model that takes the individual ply orientation and anisotropic material characteristics of a composite structure into account. Besides, further comparisons of different blade curings or the selection of other materials of both resin and actual fabric seem to be beneficial. However, a composite rotor could be successfully manufactured, compared to the initial $CAD$ rendering in Figure 10. Within the following it will now be examined whether this composite rotor does not compromises the design tool validation.

3. Results: Design Tool Validation

4. Discussion and Conclusion

Abbreviations

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