Performance evaluation and optimization of fixture adapter for oil drilling top drives

Achille Louodom Chedjou; Marnolin Querol; Xiaobo Peng; Jianren Zhou; Jaejong Park

doi:10.1515/nleng-2022-0263

Artikel Open Access

Performance evaluation and optimization of fixture adapter for oil drilling top drives

Achille Louodom Chedjou , Marnolin Querol , Xiaobo Peng , Jianren Zhou und Jaejong Park

Veröffentlicht/Copyright: 27. Februar 2023

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Abstract

A top drive is an essential mechanical device in oil field drilling since it provides the necessary torque for the drilling operations. Manufacturers in the oil and gas industry typically perform in-housing testing and classify the Safe Working Load of top drives. Testing a top drive requires a unique test stand, thus making testing top drives from other manufacturers a difficult challenge. A fixture adapter can be designed using geometric constraints and intuition to make testing apparatus semi-universal, yet they are often bulky and heavy, posing more significant safety concerns. This study aims to first numerically assess the existing fixture adapter and then structurally optimize it for enhancing its structural integrity and efficiency under various severe working environments. Therefore, finite element analysis (FEA) was performed on the existing fixture adapter, and compliance minimization topology optimization was employed. Four load and boundary conditions were used from the three most frequent operation scenarios for the fixture adapters: (i) drilling standby, (ii) staging area, (iii) drilling make-up, and (iv) break-up. The FEA results indicated that no safety factor was compromised with a 50% and 60% mass retention constraint via topology optimization compared to the original fixture adapter. The optimized fixture adapter was also tested under compression using printed 3D prototypes to validate the finite analysis and topology optimization processes.

Keywords: drilling top drives; finite element analysis; topology optimization; 3D printing; mechanical testing

1 Introduction

Oil and natural gas provide over half of the world’s energy [1]. According to the Manufacturing Energy Consumption Survey 2018, fuel use accounted for about 68% and nonfuel sources/feedstocks accounted for about 32% of the total energy use by US manufacturers in 2018 [2]. While there have been “renewable” and “sustainable” energy initiatives, they have primarily been prohibitively expensive, complex, or unreliable. The strategies of many countries are often woven around and influenced by available resources such as mineral deposits, oil, and natural gas [2]. Crude oil and natural gas are complex mixtures of hydrocarbons, non-hydrocarbons, and other trace elements. They are usually stored in the sedimentary rock of deep formations [3], sometimes even offshore, under deep waters. Oil well exploration encompasses techniques to search for hydrogen deposits, also known as geologic reservoirs, under the earth’s surface.

Oil well drilling involves cutting a circular cross-section hole into the earth to allow oil and natural gas extraction from reservoirs. The drilling for natural gas or oil wells is done following the exploration of the oil well. Today’s modern drilling rig uses top drives, which provide the torque necessary for drilling operations. The top drive can be described as a big electrical motor traveling vertically up and down in a drilling position and imparting essential torque to the drill pipe. There are several top drive manufacturers, and each manufacturer develops its own designs, which are generally classified based on their Safe Working Load as well as the type and size of the motor that provides the torque for drilling operations. For example, there are models of 175, 250, 275, 350, 500, 750, and 1,000-ton units.

Top drive manufacturers should conduct in-house testing to ensure the product meets the technical requirements per the Factory Acceptance Test and is in conformity with the American petroleum institute specs. A typical testing site is equipped with a hydraulic or electrical motor with a dynamometer to apply and measure the necessary torque during testing, which stays around 30,000 lb-ft. A duty top drive weighing up to 36,000 lb is mounted on a huge, heavy test stand. Beacuse of this high load nature of top drives, recent studies in the oil rig community have focused on patents for novel top drives [4,5], as well as the characterization [6] and modeling for performance evaluation of a specific type of top drives and its components [7].

Each manufacturer will equip the test site with an extensive test stand to perform the torque test on top drives. However, it is problematic for a manufacturer to service different configurations because there exists a design discrepancy in the mounting mechanism among top drives from different manufacturers. To perform the test on other arrangements, the manufacturer will thus have to build a big test stand for each type of configuration, which poses preventable space and financial issues. The proposed work demonstrates an attempt to design a fixture adapter that allows testing and servicing top drives from two drilling equipment manufacturers (e.g., Canrig and Tesco). The adapter should have the necessary structural stiffness to handle the stress from various configurations of top drives without excessive material usage, which can lead to potential weight-related issues. Topology optimization is used to achieve the structure that would give the minimum strain energy against the three most dominant configurations of testing apparatus (i.e., standing, staging, and operating).

Topology optimization is a newer branch of structural optimization techniques that is widely used in various industries and defense technologies. The primary benefit of topology optimization over traditional structural optimization methods, including size or shape optimization, is that it does not require a specific initial structure. Among many successful implementations, the weight of each wing was reduced by 500 kg by designing the wing box ribs using topology optimization [8]. Multimaterial structures were designed for maximum heat conductivity [9]. Topology optimization has shown its efficacy in optimizing automotive parts such as exhaust manifolds [10], knuckle joints [11], and other parts [12,13,14], to name a few. Structural optimization methods are now being applied to more complex problems, such as biomedical bone implant designs [15] and multiphysics problems [14,16,17] in larger design domains with advancements in parallel computation capabilities and efficient algorithms [18,19,20].

The initial design of the fixture adapter is proposed based on experience, intuition, and literature study, which can withstand up to 30,000 lb-ft of applied torque. The structural characteristic will be evaluated using various loadings; (i) standing: a planar component supports the upright position of the top drive during idling, (ii) staging: the fixture adapter supports the whole weight of the top drive when the assembly is not on the crane and is in the staging area, and (iii) operating: support the top drive from rotating during torque testing. The loading scenario with the highest risk of failure will be explored such that the final geometry attains a safety factor of at least 1.5. Then, the topology optimized models will be mechanically assessed in simple compression testing for proof of concept.

2 Materials and methods

The presented study starts with the proposition of an initial design obtained using SolidWorks and optimized via topology optimization with a parametric study (PS).

2.1 The initial design requirements

The fixture adapter that will accommodate the Tesco Top Drives on the Canrig test stand is designed in SolidWorks in this step. Because the system (top drive on the fixture adapter) is supposed to move vertically on the test stand via a guide runner, the design should possess mounting mechanisms to make its installation onto the test stand possible and guarantee its functionality (Figure 1a). The design should be able to receive and hold different models and configurations of top drives, especially Tesco EX300 and ES300 top drives with different mounting bases and Canrig top drive (Figure 1b and c). The adapter also needs to withstand the stresses resulting from a maximum torque of 30,000 lb-ft during torque tests.

Figure 1

Top drive test stand and various top drives that need to be accommodated: (a) Canrig test stand, (b) Tesco EX300, and (c) Canrig top drive.

To fulfill the geometric requirement for guide runners and to make the test site semi-universal, the fixture adapter in Figure 2 is proposed. It consists of four major planar supports made of 1-in thick ASTM A36 steel with a horizontal brace made of ASTM A500 high-strength steel. AISI 4140 and ASTM A106B pins were selected to secure the top drive onto the fixture.

Figure 2

(a) The initial design of the fixture adapter (yellow) is mounted on a guide runner (gray). (b) Fixture adapter with Tesco top drive is test fitted onto Canrig test stand.

2.2 Numerical modeling

The structural performance of the fixture adapter is assessed using finite element analysis (FEA) using boundary conditions and scenarios during the top drive torque test and test preparations. The goal is to ensure that the fixture adapter can sustain the loads and loading conditions that will be applied to it during torque testing.

2.2.1 System description, boundary conditions, and loads

Two significant loads are involved in this fixture adapter study: the weight of the top drive and the torque generated by the top drive during the make-up and break-up procedures. The numerical values of these parameters are provided in Table 1. It is assumed that stress values are the same regardless of the quill’s rotation direction (e.g., make-up vs break-up).

Table 1

The boundary condition for the fixture adapter

Load type	Value	Unit
The torque generated by the top drive	30,000	lb-ft
Top drive weight	36,000	lb

There are four loading scenarios exist from three distinct scenarios, as described below and in Table 2.

Table 2

Load cases and scenarios

Load case	Scenario	Description
Load case 1	Top drive in drilling standby position	The quill’s axis is vertical. The fixture adapter holds the top drive on the test stand. No torque is applied
Load case 2	Top drive in the staging area	The top drive sits on the fixture adapter. The quill’s axis is horizontal
Load case 3	Torque make-up	The crane holds top drive. The make-up torque test is in progress
Load case 4	Torque break-up	The applied torque during make-up and break-up are similar, with the only difference being the rotation of the quill is in the opposite direction of each other

Load scenarios

The planar component supports the upright position of the top drive during idling. The planar support will experience a minor bending moment due to the weight of the top drive.
The fixture adapter supports the whole weight of the top drive when the assembly is not on the crane and is in the staging area.
The fixture adapter prevents the top drive from rotating during torque testing. A primary torsion about the axis of the top drive will be applied to the fixture.

2.2.2 Forces generated in pipes and pins from make-up and break-up

The weight of the top drive will be applied straightforwardly during the study in Ansys. Yet, for the torque, due to the position of the quill or the dyno’s axis, which is far from the pipes and pins receiving the torque, it is crucial to calculate the forces resulting from the torque. The following diagrams and parameters in Figure 3 were used to compute the forces required to generate the maximum torque of 30,000 lb-ft. The resulting force values are tabulated in Table 3.

Figure 3

Load diagrams and parameters used to compute the equivalent loads for input torque. The 30,000 lb-ft of torque is translated into forces at the respective locations.

Table 3

Equivalent loads on pin and pipe for input torque during make-up and break-up

	Pin		Pipe
	F x	F y	F x	F y
lb	1248.35	754.84	1090.56	968.93

2.2.3 Meshing

Due to their acute angles, tetrahedron elements were used to mesh the eight pins. Because high-wingspan geometry preparation is unnecessary, the multizone method with an element size of 0.75″ is used on the rest of the fixture adapter, resulting in the mesh shown in Figure 4.

Figure 4

Fixture adapter design meshing in Ansys.

2.3 FEA

2.3.1 Load case 1 – drilling standby position

2.3.1.1 Loads and boundary conditions

In load case 1, the weight of the top drive is transmitted to the adapter via four pins, with the possibility of adjusting in two different positions. Since there are four pins equally loaded, each pin will support a load of 36,000/4 = 9,000 lb. Because of the possibility of adjusting in two positions, all the pins receive an equal force of 9,000 lb, or each hole receiving the pins is charged with 4,500 lb. Fixed supports are applied to the four pads where the adapter is attached to the guide runner (Figure 5); fixed supports are in purple.

Figure 5

Load and boundary conditions for load case 1.

2.3.1.2 Results

The equivalent von Mises stress plot over the entire body with the averaged display option is shown in Figure 6. The maximum value is around 23 kpsi, which occurs at both ends of the high strength steel (HSS) (8 × 6 × 3/8) and at the filled radius of 1.50 in, in the interior flat plates. This resulted in a minimum safety factor of 1.59 for the fixture adapter.

Figure 6

Equivalent (von Mises) stress and safety factor plot for load case 1 – Drilling standby position.

2.3.2 Load case 2 – staging area position

Load case 2 analyzes the situation when the assembly of the top drive on the test stand is in the staging area. This occurs at the beginning of test preparation when the top drive is placed on the fixture adapter before moving to the testing in the drilling position. In this case, the quill’s axis sits horizontally, parallel to the floor. This loading condition also appears at the end of the testing, before the top drive was removed from the test stand, when the top drive was brought back to the horizontal position. In this case, the top drive sits on six pipes, and the weight of the top drive 36,000 lb is transferred to the adapter frame via the pipes.

2.3.2.1 Loads and boundary conditions

The weight of the top drive is transmitted to the adapter via six steel pipes. Since there are six pipes, it is assumed that each pipe will hold a load of 36,000/6 = 6,000 lb, as shown in Figure 7. The fixed support is kept the same as load case 1.

Figure 7

Load and boundary conditions for load case 2.

2.3.2.2 Results

The equivalent stress (von Mises) plot showed the maximum stress value of 6081.7 psi (Figure 8, red tag). The areas of maximum stress are at both ends of the HSS (8 × 6 × 3/8) and at filled radii of R = 0.75 in, at the bottom of the big flat plates. Due to the relatively smaller stress values compared to load case 1, a minimum safety factor of 5.99 is observed, which occurs in higher-stress areas.

Figure 8

Equivalent (von Mises) stress and safety factor plot for load case 2 – Staging area position.

2.3.3 Load case 3 – torque make-up and break-up

Load case 3 analyzes the fixture adapter in a torque testing situation. The top drive is in the drilling position (the quill’s axis is vertical), aligned, and connected to the dynamometer’s axis. Here, the torque from the top drive is transferred to the fixture adapter. The fixture adapter will keep the top drive from rotating. About 30,000 lb-ft, corresponding to the torque capacity of the 500-ton top drives, was used. Most top drives have a torque capacity equal to or below that value.

2.3.3.1 Loads and boundary conditions

When the test stand is in a torque-testing scenario, it is expected that loads generated by the applied torque will get transmitted through the pipes and pins that hold the top drive through the planar structure. Thus, 30,000 lb-ft of torque is simulated via the forces computed in Section 2.2.2. The fixture adapter will sit on the guide runners, so the fixed support is at the attachment point between the fixture adapter and the guide runner. The boundary condition is provided in Figure 9.

Figure 9

Load and boundary conditions for load case 3.

2.3.3.2 Results

The von Mises stress plot over the entire geometry revealed a maximum stress value of 9216.5 psi, which occurred at the location that connects the pin to the planar support structure, as shown in Figure 10. The minimum safety factor for this load case is computed to be 5.99. This safety factor value is similar to load case 2; however, the different material properties led to a higher maximum stress than in load case 2.

Figure 10

Equivalent (Von-Mises) stress and safety factor plot for load case 3 – Torque make-up and break-up.

2.4 Structural optimization

2.4.1 Topology optimization formulation

The above static studies do not reveal any concerns regarding the numbers and plots obtained for the load cases studied. The stress contour shows that the fixture adapter is within its elastic limit for each load case scenario (see von Mises stress plots in each load case studied above). Load case 1 shows a smaller safety factor of 1.59, while the other cases offer a minimum safety factor >3. This indicates that the fixture adapter has room for improvement in structural efficiency, potentially leading to a weight reduction. In this study, traditional topology optimization for minimum compliance is used to reduce the weight of the fixture adapter while maximizing the structural stiffness using Ansys. The general multi-loading topology optimization problem has the following form:

(1) min ρ C = ∑ i = 1 N w i C i ( ρ , u ) s . t . : ∫ Ω ρ d V ≤ V * K ( ρ ) i u i = F i ,

where ρ is the density vector, C is the compliance, w is the weight factor for each load case, i represents the number of load cases, u is the displacement field, and Ω is the design domain. V * is the volume constraint and K i and F i are the stiffness and load matrix, respectively. Solid isotropic material with penalization [21,22] relaxes the problem by allowing elements to have any value between 0 and 1 with a small lower bound ( ρ min ) to avoid any singularities. Along with a parametric formula for Young’s modulus ( E ( x ) ) for elemental stiffness, the following is added to the Eq. (1).

(2) 0 < ρ min ≤ ρ ≤ 1 E ( x ) = ρ ( x ) p E 0

Here, p is the penalization factor and E 0 is Young’s modulus of the material in the solid phase.

2.4.2 Analysis settings in Ansys workbench

Figure 11 illustrates the platform for the multi-loading problem feeding the topology optimization within the Ansys workbench. The two torque scenarios (torque make-up and break-up) are similar in the fact that the rotation of the quill is in the opposite direction of each other. The i in Eq. (1) is four due to four different boundary conditions (make-up and break-up being considered separately). The penalization factor was set to three following general practice in the field [23].

Figure 11

The platform for the multi-loading problem for structural topology optimization within Ansys Workbench.

In order to promote timely convergence, the following termination conditions and analysis settings were employed.

The maximum number of iterations was set to 500;
The convergence accuracy of 0.1% was used;
The minimum normalized density ( ρ min ) was set to 0.001 to avoid any singularity problems.

2.4.3 Design domain and response constraints

In this study, the design domain is limited to the planar support structure for simplicity. Including all geometries in the optimization could result in a better overall objective (i.e., compliance). Yet, sharp corners and kinks would require significantly increased computation time. The base geometry is excluded since the proper connection to the guide runner is essential. The pins, pipes, and crossover brace bars are also excluded from topology optimization for the same reason. There is a total of four planar support structures considered as design regions in the fixture adapter, as shown in Figure 12.

Figure 12

Topology optimization design domain. Optimization regions in blue and exclusion regions in red.

The response constraint can be a mass or volume constraint obtained by setting up a percentage of mass or volume to retain in the optimized model. The response can also be defined by a stress constraint, a deformation constraint reaction force, and a natural frequency. Assuming that the material density does not change, the mass constraint is similar to the volume constraint ( V * ), in Eq. (1). The only difference is mass retention constraint defines the minimum mass of the final geometry, whereas the volume constraint ( V * ) restricts the maximum volume ratio of the resulting topology. Two different mass constraint values of 50 and 60% were explored to obtain a lighter model capable of withstanding all loading conditions without sacrificing structural requirements. Two geometric and symmetric constraints are introduced to achieve better manufacturable structures. Displacement constraints shown in Figure 12 were included to stabilize the FEA. The weights of the loads ( w i in Eq. (1)) are altered in the multi-loading formulation to specify the most critical load case during the analysis. The following table shows the PS for the given topology optimization problem.

$Figure 13 Topology optimized results from PS1 ( V * = 50 % , {V}^{* }=50 \% , equal weights): (a) density iso-line using ρ ≥ 0.6 \rho \ge 0.6 in workbench, (b) optimized geometry in SpaceClaim, and (c) computer-aided design (CAD) model created in SolidWorks.$

Figure 13

Topology optimized results from PS1 ( V * = 50 % , equal weights): (a) density iso-line using ρ ≥ 0.6 in workbench, (b) optimized geometry in SpaceClaim, and (c) computer-aided design (CAD) model created in SolidWorks.

3 Topology optimization results and design validation

Topology-optimized results from Ansys Mechanical often require postprocessing since the results usually contain sharp edges and unclear boundaries, which are not feasible for manufacturing. Thus, the optimized geometries were first identified by drawing density isolines, which were then manually processed (surface representation, i.e., stl → non-uniform rational B-spline model) in SolidWorks by tracing the isolines for further validation purposes.

3.1 Parametric study 1 (PS1)

The density isolines using ρ ≥ 0.6 for the case where the mass constraint is set to 50% and the same weight used for all load cases are shown in Figure 13a. As expected, available materials are placed near the high-stress areas, and the overall structure creates a load transfer mechanism that mimics the truss-like systems commonly obtained from prevalent topology optimization benchmark problems. The mass of the optimized fixture adapter is 1,716 lb, which is a 375.6 lb (22%) reduction over the original fixture adapter.

3.2 Parametric study 2 (PS2)

The density isolines using ρ ≥ 0.6 for the case where the mass constraint is set to 60% and the same weight used for all load cases are shown in Figure 14a. The overall geometry and the load transfer mechanism are similar to those of PS1. Still, the members connecting the supports are a little thicker (i.e., the size of the voids is smaller) due to more accessible materials. Here, overall weight reduction was measured to be 312 lb, which is equivalent to an 18.2% weight reduction. A manually processed CAD file in SolidWorks is shown in Figure 14b.

$Figure 14 Topology optimized results from PS2 ( V * = 60 % , {V}^{* }=60 \% , equal weights): (a) density iso-line using ρ ≥ 0.6 \rho \ge 0.6 in workbench and (b) CAD model created in SolidWorks.$

Figure 14

Topology optimized results from PS2 ( V * = 60 % , equal weights): (a) density iso-line using ρ ≥ 0.6 in workbench and (b) CAD model created in SolidWorks.

3.3 Parametric study 3 (PS3)

This parametric is conducted to place higher importance on the stresses associated with the torque testing (make-up and break-up). The weights on the torque testing are double than those for drilling in the standby position and staging scenarios (Table 4). The density isolines using ρ ≥ 0.6 for the case where the mass constraint is set to 50% and the same weight used for all load cases are shown in Figure 15a. The final optimized fixture adapter was measured to weigh 392 lb less than the original one, which is a 23% overall weight reduction. The lost value in PS3 is more than in PS1; however, they both satisfy the mass constraint and are thus acceptable since the mass constraint is not an equality constraint.

Table 4

Analysis settings for PS within topology optimization

Parametric study			Study 1 (PS1)	Study 2 (PS2)	Study 3 (PS3)
Mass retention constraint			50%	60%	50%
Weights (w)	Boundary condition 1	Load case 1	1	1	1
	Boundary condition 2	Load case 2	1	1	1
	Boundary condition 3	Make-up	1	1	2
	Boundary condition 3	Break-up	1	1	2

$Figure 15 Topology optimized results from PS3 ( V * = 50 % , {V}^{* }=50 \% , 1:1:2:2 weights with more emphasis on torque testing): (a) density iso-line using ρ ≥ 0.6 \rho \ge 0.6 in workbench and (b) CAD model created in SolidWorks.$

Figure 15

Topology optimized results from PS3 ( V * = 50 % , 1:1:2:2 weights with more emphasis on torque testing): (a) density iso-line using ρ ≥ 0.6 in workbench and (b) CAD model created in SolidWorks.

3.4 Numerical validation

Since optimized fixture adapter geometries are manually processed, there is a chance that the structural characteristic or optimality might have been affected. Thus, all models were subject to numerical validation using the same boundary conditions used in Section 2 with the same analysis settings. The structural responses of PS1, PS2, and PS3 results are shown in Figures 16–18 and Table 5. The dimensions of the part are provided in Figure 5.

Figure 16

Structural response of topology-optimized result from PS1 against three load cases.

Figure 17

Structural response of topology-optimized result from PS2 against three load cases.

Figure 18

Structural response of topology-optimized result from PS3 against three load cases.

Table 5

Summary of numerical validation study results (SF-safety factor)

Boundary Condition		Original	PS1	PS2	PS3
BC1	SF	1.59	1.59	1.7064	1.5832
	σ max ( psi )	23,039	39,874	32,432	36,297
	d max ( in )	0.016	0.024	0.0201	0.0243
BC2	SF	5.99	4.99	6.2226	6.2686
	σ max ( psi )	6081.7	7262.2	7024.1	6949.6
	d max ( in )	0.003	0.0039	0.0039	0.0038
BC3	SF	5.99	4.0245	4.1202	3.911
	σ max ( psi )	9216.5	20,757	20,182	23,012
	d max ( in )	0.016	0.0221	0.0209	0.0208

Topology-optimized results from the parametric results show a slight compromise in a safety factor and increased maximum stress. Yet, they are all within the elastic limit. However, it can be seen that more areas with non-zero stresses are observed against all boundary conditions, indicating better materials utilization. It is also worth noting that the PS2 result against load case 1 has a higher safety factor than the original due to the lack of sharp corners that acted as stress concentration points in the original model.

3.5 Proof of concept mechanical testing

The mechanical (physical) validation of the new fixture adapter design was conducted in a simplified environment using 3D-printed objects as proof of concept. This mechanical testing has limitations because it uses ABS plastic with vastly different mechanical characteristics from the set of metals used in the actual model. The loads applied to 3D-printed models are similar to those in load case 2 but not precisely replicate it. However, this mechanical testing is expected to show how the load transfer mechanism works in the optimized models.

The mechanical validation step will include the original and optimized models from validation study 1 and validation study 2. Validation study 3 is excluded from the testing because the safety factor obtained during the numerical validation fell below the threshold of 1.59. As a first step in creating the 3D samples, the geometries are converted into STL files, the type of file the 3D printer reads. The machine employed to make the testing samples is the Stratasys uPrint SE Plus 3D Printer. A scale of 1/12 is applied to the geometries. The material used for the samples is ABSplus™ thermoplastic. Layer thickness is set to 0.013 in (0.330 mm). The three pieces are printed simultaneously in 12 h. The finished models are shown in Figure 19.

Figure 19

3D-printed models of PS1 (left), PS2 (center), and the original fixture adapter. The parts are printed as shown.

A simple compression load with displacement control of 0.02 in/min was applied to the 3D-printed models using an Instron 5,582 testing frame. A customized fixture was used to keep the optimized models horizontal during the testing (Figure 20). The load and displacement data collected were processed in MATLAB and plotted to compare the structural responses of the different samples. Figure 21 shows all three plots combined for better comparison.

Figure 20

The simple compression testing setup on optimized and original fixture adapter. A customized fixture was used for optimized models to keep the structure as horizontal as possible during the testing.

Figure 21

Force vs displacement curve of specimens under compression tests.

The specimen of the original revised model shows a maximum load capacity of 794 lb. The model specimen from PS1 shows a mass load capacity of 430 lb, representing a 46% loss compared to the original sample. The specimen model from PS2 has a maximum load capacity of 446 lb, representing a 44% loss compared to the specimen from the original model. Comparing both optimized samples and since their losses are close to each other, it is preferable to consider the optimized model from PS1 because this represents the lighter model.

4 Conclusions

In this study, the base fixture adapter that satisfies the geometric requirements of the test stand and top drive mounting configurations was designed first. Ansys Workbench was used to perform FEA on the initial design based on load and boundary conditions commonly applied in top-drive torque tests and testing preparation, and boundary conditions were defined based on three operation scenarios. The resulting stress plots revealed the materials’ underutilization in localized regions, indicating potential targets for improvements in design and material load efficiency. The FEA-based analysis and design method has shown its efficacy on other underwater mechanical components, such as the structural performance of composite seafloor pipelines [24,25]. Multi-loading topology optimization was formulated to redistribute materials for enhanced structural stiffness and material utilization. A PS was also performed by varying response constraints (50% mass reduction and 60% mass retention on the side plates) and adding different weight factors between loading scenarios. This approach allowed us to obtain three optimized parametric models, namely, PS1, PS2, and PS3.

Subsequently, samples of optimized models were fabricated using 3D printing to validate the findings further, and then compression test experiments were performed on the optimized and 3D-printed prototype models (PS1, PS2, and PS3). A MATLAB code was written to plot the curves’ force vs displacement of the geometries. It was found that the test results of the optimized model geometries and their values of maximum load to sustain in compression were fairly close to the results from FEA numerical validation. In addition, it was found that among the three geometries of PS1, PS2, and PS3, the PS1 (50% mass constraint) showed the best alignment of validation between the mechanical testing and FEA numerical simulation while maintaining acceptable safety factor values. This study clearly showed that the topology optimization technique was an efficient tool in redesigning the existing fixture adapters, resulting in an optimal design with a considerable reduction of materials – a 50% mass reduction of side plates and a total of 22% in the entire geometry, with acceptable safety factor values. The topology optimization also resulted in reduced material costs and a weight of 170.5 kg (375.6 lb), which would make the fixture adapter easier to handle, creating a safer environment for the top drive torque testing. Future work could focus on reducing the discrepancy between the loading scenario used in the numerical simulation and the mechanical experiment to quantify the structural efficiency in real-world settings.

Funding information: The authors would like to appreciate the financial support from the National Science Foundation via award #2107140 and the Department of Energy via award DENA0003987, as well as the Research & Innovation for Scholarly Excellence grant from the Division of Research & Innovation at Prairie View A&M University.
Author contributions: A.L.C. contributed to the study’s methodology, analysis, and validation and wrote the original draft. M.Q. contributed to the study’s analysis, validation, and editing. X.P. edited the initial draft. J.Z. took part in editing, logistics, and supervision. J.P. contributed to the study’s conceptualization, methodology, validation, supervision, and edited the draft.
Conflict of interest: The authors state that there is no conflict of interest.

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Received: 2022-09-24

Revised: 2022-11-14

Accepted: 2022-12-02

Published Online: 2023-02-27

This work is licensed under the Creative Commons Attribution 4.0 International License.

Artikel in diesem Heft

https://doi.org/10.1515/nleng-2022-0263

Schlagwörter für diesen Artikel

drilling top drives; finite element analysis; topology optimization; 3D printing; mechanical testing

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