Performance Analysis of Natural Gas Centrifugal Compressors Under Hydrogen-Blended Conditions

Abstract

The transport of natural gas blended with hydrogen is a key strategy for the low-carbon energy transition. However, the influence mechanism of its thermo-physical property variations on centrifugal compressor performance remains insufficiently understood. This study systematically investigates the effects of the hydrogen blending ratio (HBR, 0–30%), inlet temperature, and rotational speed on key compressor parameters (pressure ratio, polytropic efficiency, and outlet temperature) through numerical simulations. In order to evaluate the influence of hydrogen blending on the performance and stability of centrifugal compressors, a three-dimensional model of the compressor was established, and the simulation conducted was verified with the experimental data. Results indicate that under constant inlet conditions, both the pressure ratio and outlet temperature decrease with increasing HBR, while polytropic efficiency remains relatively stable. Hydrogen blending significantly expands the surge margin, shifting both surge and choke lines downward, and consequently reducing the stable operating range by 27.11% when hydrogen content increases from 0% to 30%. This research provides theoretical foundations and practical guidance for optimizing hydrogen-blended natural gas centrifugal compressor design and operational control.

1. Introduction

Under the dual drivers of global energy transition and climate governance, innovative breakthroughs in green energy technologies have become critical in addressing energy crises and environmental challenges. As a core component of clean energy systems, hydrogen energy is internationally recognized as the most promising strategic energy source due to its sustainable resource availability, high energy density advantages, and zero-pollution characteristics [1,2]. However, the commercial application of hydrogen energy still faces practical challenges such as high storage/transportation costs, insufficient infrastructure, and prominent safety risks [3]. In this context, natural gas–hydrogen blending technology by injecting hydrogen into existing natural gas pipelines not only enables economical and efficient hydrogen transportation but also optimizes natural gas combustion characteristics, significantly reducing carbon emission intensity. Globally, academic institutions, industry, and governments are supporting hydrogen blending projects—like HyDeploy, GRHYD, THyGA, and HyBlend—all aimed at developing efficient pathways to meet carbon reduction goals in the coming decades [4]. This approach provides an innovative solution for global low-carbon energy transition [5,6,7].

With the rapid development of hydrogen blending technology, research on the physical property evolution mechanisms and transport characteristics of mixed gases has emerged as a developing topic in international academia and engineering [8]. Fausto Arpino et al. conducted research demonstrating that distributed green hydrogen production via photovoltaic-powered electrolysis, integrated with Italy’s natural gas transmission network (NGTN), can achieve 0–0.90 vol% hydrogen blending with optimal storage, reducing production costs to 5.10 EUR/kg H₂ and saving 32.348 tons of CO₂ annually per 500 m² PV installation, thereby addressing energy storage challenges for intermittent renewables while decarbonizing existing infrastructure [9]. Studies reveal that hydrogen blending substantially alters the fundamental physical parameters of natural gas, thereby influencing the operational characteristics of entire pipeline systems [10]. Notably, the performance variations in centrifugal compressors as critical power equipment are particularly significant. Among existing research methods, similarity theory has been widely adopted for hydrogen blending impact assessments due to its simplicity [11,12,13]. Q. Peng et al. [14] found that a Gray-RBF neural network prediction model improved the prediction accuracy of natural gas blended with hydrogen. L. Wang et al. [15] propose a similarity transformation method for centrifugal compressors based on the predictor-corrector method to predict compressor performance parameters. P. Wang et al. [16] proposed predictive correction algorithms to enhance pressure ratio conversion accuracy. Nevertheless, these methods rely on ideal gas assumptions and simplified models, resulting in inherent limitations in prediction precision. Current research exhibits notable gaps in understanding the coupled influence between hydrogen blending ratios and operational parameters, necessitating deeper exploration.

Numerical simulation technology, leveraging its advantages of high precision and efficiency, is increasingly becoming the mainstream approach for compressor performance research [17,18]. Elias et al. [19] uncovered compressor stall mechanisms through Large Eddy Simulations; Gu et al. [20] analyzed thermal conduction effects using the SST turbulence model; and Xiong et al. [21] evaluated hydrogen blending impacts on operational stability via the SA turbulence model. Notably, recent studies demonstrate that compressor stability can be enhanced through surface engineering—Khan et al. proved that applying 200 μm of engineered roughness on the diffuser shroud expands the stable operating range by 18% and stall margin by 7.98% by suppressing flow separation, albeit with a trade-off in pressure ratio and entropy generation [22].

However, despite these advancements, several critical gaps remain when focusing specifically on centrifugal compressors for hydrogen-blended natural gas (HBNG). First, many predictive models rely on similarity theory, which often incorporates ideal gas assumptions, potentially leading to inaccuracies for real-gas HBNG mixtures under high pressure. Second, existing numerical studies frequently focus on individual performance parameters (e.g., pressure ratio) but lack a systematic analysis of the coupled influence of HBR with other operational variables like rotational speed and inlet conditions. Third, and most importantly, there is a scarcity of quantitative data on how HBR affects the entire stable operating range of the compressor, particularly the shift in surge and choke boundaries, which is paramount for operational safety and system flexibility.

This study establishes a three-dimensional numerical model of a centrifugal compressor to systematically investigate the interactive effects of multiple factors, including rotational speed, hydrogen blending ratio, and inlet parameters, on compressor performance. Through systematic analysis, we provide a systematic investigation of the coupled effects of hydrogen blending ratio (0–30%), rotational speed, and inlet conditions on the performance of a two-stage centrifugal compressor (Rolls-Royce Energy compressor, London, UK), moving beyond single-factor analyses. Unlike previous studies focusing primarily on efficiency or pressure ratios, we quantitatively demonstrate how hydrogen blending significantly narrows the compressor’s stable operating range by shifting the surge and choke boundaries, providing crucial data for operational risk assessment. This study provides theoretical foundations and technical support for optimizing the design and ensuring safe operation of hydrogen-blended systems.

2. Methods

2.1. Model Establishment

This study focuses on a representative two-stage centrifugal compressor characterized by a first-stage impeller diameter of 993.77 mm (Rolls-Royce, London, UK), operating at a rated speed of 4800 rpm and a power of 40 kW. The compressor has a design-point flow rate of 5.3 m³/s under specified inlet conditions (6 MPa, 285 K), achieving an overall pressure ratio of 1.6. Based on the basic size parameters of the centrifugal compressor (Table S1), the impeller of the compressor was modeled using the CFTurbo 10.3 software, and the three-dimensional models of the two-stage impellers of the compressor were established, respectively (Figure S1). The data fluid domain model required for the impeller was extracted through Boolean operations, and the fluid domain was redefined using structural data in SpaceClaim 2023. To prevent the backflow of the outlet gas, we have also set up an extension section at the outlet. Subsequently, the fluid domain is subjected to grid division to form a structural grid. The automatic grid division module is used to achieve the matching of the grid division size and structure. The orthogonality quality of the grid is above 0.2. Different grid parts that meet the calculation requirements are shown in Figure S1.

2.2. Governing Equations and Turbulence Models

When natural gas mixed with hydrogen flows through a centrifugal compressor, its motion follows the conservation laws of mass, momentum, and energy. Based on these physical conservation laws, the continuity equation, momentum conservation equation, and energy conservation equation for the gas mixture can be derived. The continuity equation is established according to the mass conservation law and can be expressed in the Cartesian coordinate system as follows:

$${\displaystyle \frac{\partial \rho }{\partial t}}+div\left(\rho u\right)=0$$

(1)

The momentum equation [23] is derived from Newton’s second law of motion. In the Cartesian coordinate system, the momentum equations in the X, Y, and Z directions can be expressed as follows:

$${\displaystyle \frac{\partial \left(\rho u_x\right)}{\partial t}}+div\left(\rho u_{x}u\right)=F_x-{\displaystyle \frac{\partial \rho }{\partial x}}+{\displaystyle \frac{\partial {\tau }_{xx}}{\partial x}}+{\displaystyle \frac{\partial {\tau }_{yx}}{\partial y}}+{\displaystyle \frac{\partial {\tau }_{zx}}{\partial z}}$$

(2)

$${\displaystyle \frac{\partial \left(\rho u_y\right)}{\partial t}}+div\left(\rho u_{y}u\right)=F_y-{\displaystyle \frac{\partial \rho }{\partial y}}+{\displaystyle \frac{\partial {\tau }_{xy}}{\partial x}}+{\displaystyle \frac{\partial {\tau }_{yy}}{\partial y}}+{\displaystyle \frac{\partial {\tau }_{zy}}{\partial z}}$$

(3)

$${\displaystyle \frac{\partial \left(\rho u_z\right)}{\partial t}}+div\left(\rho u_{z}u\right)=F_z-{\displaystyle \frac{\partial \rho }{\partial z}}+{\displaystyle \frac{\partial {\tau }_{xz}}{\partial x}}+{\displaystyle \frac{\partial {\tau }_{yz}}{\partial y}}+{\displaystyle \frac{\partial {\tau }_{zz}}{\partial z}}$$

(4)

The momentum equations are the Navier–Stokes (N-S) equations, where the terms ${\tau }_{xx}$, ${\tau }_{xy}$ and ${\tau }_{xz}$ represent the surface viscous stresses generated by molecular interactions. F_x, F_y, and F_z denote the body forces. Since the compressor model only considers the influence of gravity, these terms are defined as F_x = 0, F_y = 0, and F_z = −ρg. The energy equation can be expressed as follows:

$${\displaystyle \frac{D}{Dt}}{e}_s=F_{b1}v+{\displaystyle \frac{1}{\rho }}div\left(P·v\right)+{\displaystyle \frac{1}{\rho }}div\left(kgradT\right)+q$$

(5)

where e represents the specific internal energy (per unit mass) due to random molecular motion, ρ denotes the fluid density, t is time, v is the fluid velocity vector, and P stands for the pressure tensor, which characterizes the stress acting on any portion of the fluid element, including both normal and shear stresses.

The Spalart–Allmaras (S-A) [24,25] turbulence model was selected based on computational accuracy considerations. This model accounts for boundary viscosity effects and closely approximates actual fluid flow behavior, capable of providing accurate predictions for the flow phenomena of interest. Furthermore, the S-A model offers rapid convergence rates, relatively lenient near-wall grid resolution requirements, and lower computational memory demands, representing an optimal balance between solution speed, accuracy, and stability. Therefore, this study employs the S-A turbulence model to simulate the internal flow field in a two-stage centrifugal compressor [24]. Although S-A is a reasonable choice and has been confirmed in some studies that its prediction results are closer to the experimental results, for some analyses, the correct capture of the flow field structure is more important than the performance parameters of the compressor. More advanced two-path models, such as SST k − ω, are also suitable for turbo mechanical applications. The transport equation for the S-A model is expressed as follows:

$${\displaystyle \frac{\partial \tilde{v}}{\partial t}}+\overrightarrow{V}·\mathsf{\nabla }\tilde{v}={\displaystyle \frac{1}{\sigma }}\left\{\mathsf{\nabla }·\left[\left(v+\left(1+c_{b2}\right)\tilde{v}\right)\mathsf{\nabla }\tilde{v}\right]-C_{b2}\tilde{v}\mathsf{\nabla }\tilde{v}\right\}+Q$$

(6)

In the equation, σ and Cb₂ are model constants, $\overrightarrow{V}$ represents the mean velocity vector, and Q denotes the source term. The source term Q consists of production and dissipation terms. Additionally, the constants in the S-A turbulence model are given by the following expressions:

$$C_{w1}=C_{b1}+{\kappa }^2{\displaystyle \frac{1+C_{b1}}{{\sigma }_v}}$$

(7)

where σ_v is the turbulent diffusion coefficient, $\kappa$ denotes the von Kármán constant, and Cb₁ represents a model constant.

2.3. Boundary Condition Settings

In the simulation, the thermophysical properties of hydrogen-blended natural gas were determined based on the GERG-2008 equation of state [26], Lohrenz–Bray–Clark viscosity equation, the thermal conductivity model, and the volumetric energy density equation [27]. Detailed parameter values are provided in the Supplementary Materials. The study focuses on the performance of the compressor under steady conditions. The steady-state simulation method offers significant advantages in computational efficiency, making it well-suited for the extensive calculations required in performance mapping. Within the stable operating range, the simulation results agree well with experimental data [28], a consistency further validated in a subsequent section of this paper. Given that the core objective of this research is to analyze the influence of different hydrogen blending ratios, the conclusions derived from this method are sufficient to provide qualitative guidance for future studies.

Inlet Boundary Conditions: A pressure inlet boundary was specified, with both inlet pressure and temperature set according to simulation requirements. The boundary conditions can be expressed as follows:

P_in = constant, T_in = constant

(8)

where P_in and T_in represent the total pressure (Pa) and temperature (K) at the compressor inlet, respectively.

Outlet Boundary Conditions: A mass-flow outlet boundary was specified. The mass flow rate was determined by the inlet composition, hydrogen blending ratio, inlet pressure, and inlet volumetric flow rate. The specific mass flow value was input according to the simulation conditions. The boundary condition can be expressed as follows:

ρ_outV_out = constant

(9)

In the equation, ρ_out and V_out denote the outlet fluid density (kg/m³) and volumetric flow rate (m³), respectively. The interface between the stationary and rotating components was treated using a stage (mix-plane) model.

Impeller Rotation Speed: Ten discrete rotational speeds were simulated within the range of 3120–5040 rpm. The outlet parameters of the first-stage compressor were directly assigned as the inlet conditions for the second stage, with the final results representing the combined output of the two-stage compression process.

Wall Boundary Conditions: Standard wall functions were applied with default surface roughness and a no-slip boundary condition. Mathematically, this can be expressed as follows:

V_wall = 0, ∂T/∂n = 0

(10)

V_wall and ∂T/∂n represent the wall velocity and the wall temperature gradient, respectively. Since the compressibility of the fluid cannot be ignored in the simulation process of the two-stage centrifugal compressor, a pressure-based solver is adopted, the energy equation is enabled, and the explicit coupling algorithm of pressure–velocity coupling is selected. In the CFD simulation (Ansys CFX 2023) of the centrifugal compressor, the residual convergence criterion was set to 10⁻⁶, and the total outlet pressure, outlet temperature, enthalpy change and the torque acting on the impeller were monitored simultaneously.

The pressure ratio is the ratio of the discharge pressure to the inlet pressure of a gas, representing the increase in gas head and serving as a key performance indicator for compressors. The general equation for pressure ratio is as follows:

$$\epsilon ={\displaystyle \frac{P_{out}}{P_{in}}}$$

(11)

ε is the pressure ratio of the compressor and P_out is the outlet pressure (Pa).

The compressor polytropic efficiency η is defined as follows:

$$\eta =\left({\displaystyle \frac{k-1}{k}}\right){\displaystyle \frac{\mathit{lg}\left(\frac{P_{out}}{P_{in}}\right)}{\mathit{lg}\left(\frac{T_{out}}{T_{in}}\right)}}$$

(12)

where T_out is the outlet temperature (K). The k is the isentropic exponent.

The power P (W) of a compressor can be calculated from the torque M (N·m) and rotational speed n (rad/s) using the following equation:

$$P=M\ast n$$

(13)

2.4. Model Validation

To ensure the numerical simulation results are grid-independent [22], grid independence verification was performed using pure natural gas as the working fluid. The impeller section is densified by the boundary layer, keeping the Y+ value close to 1 to ensure the applicability of the turbulence model. The computational domain was discretized with varying grid cell quantities under boundary conditions of 285 K inlet temperature, 6 MPa inlet pressure, 197 kg/s outlet mass flow rate, and 4800 rpm impeller speed. As shown in Figure 1a, the influence of grid size becomes negligible when the cell count exceeds 700,000, with total pressure ratio remaining essentially constant and outlet temperature variations below 1%. Considering both computational accuracy and efficiency, a grid configuration of 800,000 cells per impeller was selected for subsequent simulations.

A comparative study between experimental measurements and simulation results was conducted to validate the numerical model’s reliability (Figure 1b and Figure S2 and Table S4). The experimental data are derived from the original data of the actual on-site operation of the West–East Gas Pipeline compression station in China. Some of the actual operational data of the compressor is shown in Figure S3. Figure 1b demonstrates that the maximum deviation of total pressure ratio is approximately 3.05%, while the discrepancy in outlet temperature remains within 1.95%. These margins fall within acceptable engineering tolerances, particularly considering complex flow phenomena such as shock wave interactions in compressible fluids and secondary flows within impeller passages. The observed agreement confirms the rationality of boundary condition settings and gas state equation implementations in the numerical model. This validated approach establishes a robust foundation for subsequent hydrogen-blending optimization and compressor redesign studies.

3. Results and Discussion

This section conducts an in-depth analysis of how varying hydrogen blending ratios affect compressor performance under different operating conditions, with a focus on key compressor performance parameters and operational ranges.

3.1. Research on Operating Characteristics of Centrifugal Compressors Under Pure Natural Gas Conditions

We analyzed the operating characteristic curves and operational ranges of the compressor under pure natural gas compression conditions at different rotational speeds, with the pressure ratio curve, outlet temperature curve, power curve, and polytropic efficiency curve shown in Figure 2. The results demonstrate that the compressor has both maximum and minimum operating speed limits, with each speed corresponding to specific maximum and minimum inlet flow boundaries. As seen in Figure 2a, the total pressure ratio shows a clear decreasing trend as speed decreases. When the speed decreases from 5040 r/min to 3120 r/min, the maximum pressure ratio drops from approximately 1.79 to 1.27, directly reflecting the reduced gas kinetic energy conversion efficiency caused by decreased impeller linear velocity. Simultaneously, the stable operating flow range of the compressor narrows, manifested by both the surge boundary and choke boundary shifting toward lower flow rates. This occurs because gas flow stability deteriorates at lower speeds, making flow separation more likely to occur at the impeller inlet. With constant inlet temperature, the outlet temperature decreases monotonically as compressor speed reduces, showing a total drop of about 30 K from maximum to minimum speed, primarily attributed to the reduced mechanical energy obtained by the gas during compression (Figure 2b). The shaft power decreases significantly with reduced speed, consistent with the similarity laws of centrifugal compressors (Figure 2c). It is worth noting that the overall polytropic efficiency of the compressor varies slightly at different rotational speeds, indicating that the compressor maintains relatively good aerodynamic performance over a wide speed range (Figure 2d). This characteristic is particularly important for stable operation under variable working conditions, especially in future hydrogen-blending applications requiring frequent load adjustments.

Through systematic investigation of inlet parameter effects on centrifugal compressor performance, the characteristic curves shown in Figure 3 were obtained. When the inlet temperature increased (285 K–315 K), the 4% reduction in compressor pressure ratio can be attributed to the increased gas specific heat capacity at higher temperatures, which reduces the compression work per unit mass gas at constant rotational speed (Figure 3a). Simultaneously, according to the ideal gas law, elevated temperature decreases gas density and increases volumetric flow, shifting the compressor operating point toward higher flow rates—another key factor for the pressure ratio decline. The 38 K rise in outlet temperature directly obeys the first law of thermodynamics, where higher initial enthalpy leads to correspondingly elevated final temperature after near-adiabatic compression (Figure 3b). Simulation results demonstrate that pressure ratio and outlet temperature exhibit only minor fluctuations during inlet pressure variations, indicating the compressor’s strong adaptability to pressure changes under constant speed, temperature and flow conditions (Figure 3c,d).

3.2. Research on the Operational Characteristics of Centrifugal Compressors Under Hydrogen-Blended Conditions

3.2.1. Impact of Hydrogen Blending on Compressor Pressure Ratio

Building upon our research on the characteristic curves of natural gas compression in centrifugal compressors without hydrogen blending, we further investigated the performance curve variations under different hydrogen mixing ratios. As shown in Figure 4 and Figure S2, the performance curves of the centrifugal compressor shift downward and to the left with increasing hydrogen blending ratios. The leftward shift indicates reduced flow capacity at equivalent pressure ratios, while the downward shift reflects pressure ratio deterioration at identical flow rates. When the hydrogen blending ratio reaches 30%, the pressure ratio decreases by an average of approximately 13.5%. This phenomenon occurs because hydrogen has a lower density than natural gas, and hydrogen–natural gas mixtures reduce the overall density of pipeline-transported gas. Under identical pressure conditions, hydrogen exhibits a higher compressibility factor than natural gas, making it more difficult to compress. Consequently, the same volume of mixed gas occupies more space, requiring greater velocity and power for transportation. However, with the compressor’s rotational speed remaining constant, the centrifugal compressor can only adapt to the changes in pipeline medium by reducing the inlet flow rate, resulting in a downward shift in the performance curve. The higher the hydrogen proportion, the more pronounced this downward shift becomes. At flow rates of 4.3 m³/s, 5.6 m³/s, and 7.0 m³/s, the pressure ratios decrease by 14.8%, 13.7%, and 12.5%, respectively, when the hydrogen blending ratio reaches 30%. The impact of hydrogen blending on the pressure ratio gradually diminishes as the flow rate increases. Under low-flow conditions, where gas kinetic energy and velocity are lower, the changes in density, specific heat capacity, and temperature induced by higher hydrogen blending ratios have a more significant effect on the compressor’s compression efficiency and overall pressure ratio.

Figure 5 shows the total pressure ratio of the compressor under different inlet flow rates and hydrogen blending ratios. When the impeller speed remains constant, the variation trend of the total pressure ratio with hydrogen blending ratio (from 0% to 30%) is consistent across different flow rates, but the magnitude of change differs. At an impeller speed of 3600 r/min, the total pressure ratio decreases by an average of 10.7% when the hydrogen blending ratio increases from 0% to 30%; at 4330 r/min, the average reduction is 12.2%; and at 4800 r/min, the average reduction reaches 13.5%. This demonstrates that higher rotational speeds amplify hydrogen’s impact on the total pressure ratio. This speed dependency arises because increased gas kinetic energy at higher speeds further accentuates the influence of hydrogen’s low-density characteristics on the flow dynamics.

3.2.2. Impact of Hydrogen Blending on Compressor Outlet Temperature

Through a comprehensive investigation of compressor outlet temperature characteristics, we have identified that hydrogen blending ratios influence thermal behavior in distinctive patterns (Figure 6 and Figure 7). Under varying flow conditions, the outlet temperature variation remains within a modest 5 K range when hydrogen content increases from 0% to 30%. While hydrogen possesses a higher specific heat capacity than natural gas, the gradual evolution of mixed gas properties and the predominant dependence of temperature rise on compression work and thermodynamic processes collectively constrain outlet temperature fluctuations. Particularly significant is the minimal temperature variation (under 1 K) observed across different rotational speeds (3120–5040 rpm) with hydrogen blending. These findings demonstrate the relatively low sensitivity of compressor outlet temperature to hydrogen blending ratios compared to critical parameters like pressure ratio, offering valuable engineering implications. This characteristic enables operational focus to shift from temperature monitoring to more responsive parameters, such as pressure ratio during hydrogen proportion adjustments.

3.2.3. The Impact of Hydrogen Blending on Pressure Ratio and Outlet Temperature

This study analyzes the impacts of hydrogen blending (0–30%) on a centrifugal compressor operating at 285 K inlet temperature, 6 MPa inlet pressure, and 4800 r/min impeller speed. Figure 8 shows that with fixed speed and inlet pressure, rising inlet flow rate reduces both total pressure ratio and outlet temperature, while higher hydrogen blending ratios at constant volumetric flow further decrease these parameters due to reduced gas density and increased mixture-specific heat capacity. Figure S5 extends these findings across varying inlet pressures.

Figure 6. Variation in outlet temperature with different hydrogen blending ratios at 4800 r/min.

Figure 6. Variation in outlet temperature with different hydrogen blending ratios at 4800 r/min.

Figure 7. Flow rate–outlet temperature characteristics of the compressor at different hydrogen blending ratios. The curves depict the compressor’s outlet temperature variations across flow rates under hydrogen blending ratios of (a) 5%, (b) 10%, (c) 15%, (d) 20%, (e) 25%, and (f) 30%.

Figure 7. Flow rate–outlet temperature characteristics of the compressor at different hydrogen blending ratios. The curves depict the compressor’s outlet temperature variations across flow rates under hydrogen blending ratios of (a) 5%, (b) 10%, (c) 15%, (d) 20%, (e) 25%, and (f) 30%.

3.2.4. The Impact of Hydrogen Blending on Compressor Power and Polytropic Efficiency

Figure 9 illustrates the influence of hydrogen blending ratio on compressor shaft power at different rotational speeds. Under the fixed-speed condition of 4800 r/min, as the hydrogen blending ratio increases from 0% to 30%, the shaft power characteristic curve exhibits a clear downward shift, with an overall reduction of approximately 30%. This phenomenon can be reasonably explained by the shaft power calculation formula. First, the density of hydrogen is significantly lower than that of natural gas, leading to a notable decrease in the mixed gas density as the hydrogen blending ratio rises. Second, the energy head, being a function of speed and flow rate, undergoes changes due to flow redistribution at constant speed. Finally, although efficiency may fluctuate due to variations in gas properties, the reduction in density remains the dominant factor in the power decrease.

We further reveal the influence of rotational speed on shaft power characteristics (Figure 10). When maintaining the same hydrogen blending ratio, the shaft power increases significantly with higher rotational speeds. The power reduction caused by hydrogen blending shows slight variations at different speeds, with the decrease being slightly smaller (by about 5%) in the low-speed range compared to the high-speed range. At higher speeds, the increased kinetic energy of the gas amplifies the impact of density changes on power. Additionally, the varying flow rate ranges under different speed conditions lead to differences in the contribution of energy head to power.

Figure 11 demonstrates the variation patterns of centrifugal compressor polytropic efficiency characteristics under different rotational speeds and flow conditions. The research results indicate that compressor polytropic efficiency exhibits a typical parabolic trend with flow rate variation reaching peak polytropic efficiency near the design flow rate, then gradually decreasing as the flow rate either increases or decreases further. At the design flow condition, the aerodynamic matching between the impeller and diffuser is optimal, with minimal flow separation and vortex losses. However, when operating away from the design flow rate, secondary flows intensify within the flow passages, leading to increased losses. Notably, the influence of hydrogen blending ratio on polytropic efficiency is negligible (variation amplitude < 1%). This is primarily because polytropic efficiency, as a dimensionless parameter characterizing energy conversion effectiveness, mainly depends on the geometric features of flow components and flow conditions, while being less sensitive to medium properties. Further analysis reveals that, under a fixed hydrogen blending ratio, the increasing rotational speed shifts the polytropic efficiency curve toward higher flow rates. The flow coefficient and specific speed jointly determine performance curve migration. Higher rotational speeds require proportionally increased volumetric flow to maintain a constant flow coefficient at homologous operating points, thereby shifting the best polytropic efficiency point toward higher flow regimes.

3.3. Analysis of the Impact of Different Hydrogen Blending Ratios on Compressor Stability Operating Range

The centrifugal compressor operates at different speeds, and the variation in stable operating ranges under different hydrogen blending ratios is shown in Figure 12. The maximum speed line, minimum speed line, surge boundary line, and maximum flow line collectively define the stable operating range of the centrifugal compressor. The surge line was determined using a common approximation method. At a given rotational speed, a series of simulations was performed by gradually reducing the mass flow rate. The point of maximum pressure ratio on the performance curve was defined as the surge boundary. The compressor must operate within this range to ensure safety. When hydrogen is mixed with natural gas, differences in thermophysical properties affect the compressor’s characteristic curves. This section analyzes the impact of hydrogen on the operating range of the centrifugal compressor.

Figure 12 illustrates the compressor’s operating range when delivering natural gas at an inlet temperature of 275 K and an inlet pressure of 6 MPa. Numerical calculations were used to determine the operating range under different hydrogen blending ratios. In the figure, the line between points A–D forms the maximum speed line; the line between points B–C forms the minimum speed line; the line between points A–B forms the surge boundary line; and the line between points D–C forms the maximum flow line. It can be observed that as hydrogen blending increases, the compressor’s stable operating range gradually narrows.

From the changes in the compressor’s operating range characteristic curves under different blending ratios in Figure 12, it is evident that when the hydrogen proportion increases from 0% to 30%, the maximum speed line shifts downward, making the compressor more susceptible to aerodynamic instability during high-speed operation. Using the polygon area module in ImageJ 2025 software to calculate the enclosed area of the six curves, the results show that as hydrogen blending increases to 30%, the total operating area decreases by 4.28%, 9.11%, 13.24%, 17.32%, 22.81%, and 27.11% compared to the 0% blending ratio.

Figure 13a demonstrates the variation trend of the surge boundary line after green hydrogen blending. As the blending ratio increases, the surge line shifts downward, expanding the surge region and reducing the stable operating range. This phenomenon primarily results from hydrogen’s lower density, which decreases the overall density of the gas mixture, alters momentum and pressure characteristics, and simultaneously intensifies flow separation and vortex formation. The expanded surge range will increase operational risks for centrifugal compressors. Figure 13b illustrates the variation in the maximum flow line after blending. With increasing hydrogen proportion, the maximum flow line moves downward, indicating reduced flow delivery capacity under identical operating conditions, which stems from deteriorated aerodynamic performance. Figure 13c,d present the variation patterns of the minimum and maximum speed lines with increasing hydrogen blending ratios. Both speed lines shift downward, leading to a significant reduction in the total pressure ratio. This is mainly attributed to hydrogen’s low-density characteristics affecting the compressor’s ability to generate sufficient pressure rise.

The hydrogen blending in compressor operation significantly increases the probability of entering the surge region, as shown in Figure 14. When the compressor’s rotational speed and pressure ratio remain constant, point A during pure natural gas transportation shows an inlet flow rate of approximately 5.7 m³/s, which remains far from the surge region. However, at point B with a 20% hydrogen–natural gas blend, the flow rate decreases to 3.6 m³/s and precisely coincides with the surge point. Consequently, the reduced operating flow rate after hydrogen blending implies an increased likelihood of the compressor’s working point entering the surge region, thereby diminishing the pipeline system’s adjustability. In practical operation, measures should be taken to regulate the compressor’s flow rate to maintain a safe distance from the surge region.

In fact, these quantified performance changes and the contraction of the stable operating range provide actionable guidelines for the deployment of hydrogen and natural gas infrastructure. The selection of compressors must take into account the power derating caused by hydrogen, and the capacity margin of the unit should be increased. The operation plan should take into account real-time speed adjustment and flow control to counteract the surge boundary displacement. The safety monitoring system needs to recalibrate the stability threshold related to H₂, especially when the mixing ratio is high, as the risk of a surge will increase.

4. Conclusions

This study employed numerical simulation to analyze the performance of a centrifugal compressor in transporting hydrogen-blended natural gas. The feasibility of the methodology was confirmed by comparing simulation results with experimental data for pure natural gas transportation. Centrifugal compressors exhibited significant aerodynamic performance variations under hydrogen-blended conditions. As the hydrogen blending ratio increased from 0% to 30%, the compressor pressure ratio decreased by 10–13.5% on average, the shaft power reduced by approximately 30%, and polytropic efficiency remained nearly unchanged with less than 1% variation. Concurrently, the outlet temperature decreases by approximately 5 K, indicating robust operational adaptability. Notably, hydrogen blending substantially narrows the stable operating range by up to 27.1%, primarily due to a downward shift in the surge line that increases operational risks. To ensure stable compressor operation while meeting compression requirements for hydrogen–natural gas transportation, we recommend appropriately increasing the inlet mass flow rate and moderately elevating rotational speed.