Mixed Flow Pumps

As with any turbomachinery, pump design requires a lot of effort on finding the right blade profile for the specified application. As there is no right or wrong in the process, engineers have to make some general assumptions as a starting point. Generally, we can say that the focus of this task is to minimize losses. It is obvious that the selected blade shape will affect several important hydrodynamic parameters of the pump and especially the position of optimal flow rate and the shape of the overall pump performance curves. In addition to axial and radial pump design in recent years, we also have seen the development of mixed-flow pumps. A mixed flow pump is a centrifugal pump with a mixed flow impeller (also called diagonal impeller), and their application range covers the transition gap between radial flow pumps and axial flow pumps.

Let’s consider a dimensionless coefficient called “specific speed” in order to be able to compare different pumps with various configurations and features. The “specific speed” is obtained as the theoretical rotational speed at which a geometrically-similar impeller would run if it were of such a size as to produce 1 m of head at a 1l/s flow rate. In formulas:
formulawhere ns is the specific speed, n the rotational speed, Q is the volume flow rate, H is total head and g is gravity acceleration.

Fig1: Picture courtesy of KSB Pumps

The specific speed for mixed flow pumps might go from ns=30  to ns=80  for low-speed mixed flow pumps and can increase between ns of 80 and 160  for higher speed mixed flow pumps, as shown in Fig 1 to the left.

As you can see, Mixed-flow pumps have specifics speeds right in-between radial and axial configurations. For mixed flow pumps we can mainly see two casing configurations. In case of low peed pumps, the most appropriate configuration is the volute casing, which shows big similarities with centrifugal pump designs. However, for higher specific speeds raising above the ns =~130/140  value, we see configurations combining the  mixed flow impeller with a diffuser + tubular casing. This solution is chosen to avoid the unreasonably large volute outlet cross sections that would be required to maintain the flow due to the very low tangential component of the absolute velocity at the impeller outlet.

Fig2: Picture courtesy of Sulzer Pumps

Mixed-flow pumps with tubular casing are often installed in a vertical arrangement, and can be found in multistage configuration for high flow applications, such as irrigation, urban water supply, thermal power plants etc.

In a nutshell, the advantage of mixed-flow pumps over the “competition” is to be a “jack of all trades”. They combine higher massflow of axial configurations with higher pressure achievable by centrifugal machines.


Foundations of Rotordynamics – Part 1

In order to succeed as an engineer focused on rotor dynamics in rotating equipment, it is important to be fully aware of its foundation. The foundations of rotor dynamics consists of two parts, lateral analysis and torsional analysis. For part one of this blog, I will be focusing on lateral analysis and then exploring torsional analysis in part two next week.

Lateral analysis, also referred as critical speed analysis, is the study of when rotational speed meets or exceeds the shaft natural frequency. This is important since not knowing the critical speed can lead to instability, unbalance or even cause unknown forces to alter the functionality of rotating machinery. Since a rotating machinery consists of many components (rotor, bearing, motor, seals, etc), lateral analysis is made up of three categories: undamped critical speed analysis, steady state synchronous analysis (also known as damped unbalance response analysis) and stability analysis.

Influenced by the rotor’s mass and stiffness properties, undamped critical speed analysis is used for the estimation of critical speeds, mode shape characteristics and eigenvalues. Generally, this analysis excludes any damping in the system as well as any unbalance forces. For the estimation of undamped critical speed, an undamped critical speed map is a common tool that can be used. This map generally represents the first four undamped, forward-whirling modes as a total bearing/support stiffness. Calculated using bearing principles stiffness, mode shapes are helpful because they provide an approximate indication of the relative displacements that the shaft undergoes when the rotor operates in the vicinity of the associated critical speeds.

Figure 1: Example of critical speed calculation on front bearing using AxSTREAM RotorDynamics

Next, unbalance response analysis takes into consideration all damping effects. There are three requirements needed to satisfy unbalance performance: separation between critical speeds and operational speeds, operational speed should not be exceeded, and no rubbing should occurs when the rotor’s balance stage degrades to the probe vibration limit. Performing this analysis confirms whether or not these requirements have been achieved. It is a good idea to keep your API requirements handy when evaluating this analysis for reference.

Latestly, stability analysis calculates damped eigenvalues, and also takes into consideration oil whirl and shaft whip to avoid self-exited instabilities. It is known, that if the damping exponent of an eigenvalue is negative, then stable rotor vibrations occur, and if positive, the oppsite effect will occur. Understanding these categories relating back to lateral analysis as a whole, will allow the respective engineer, to design and analyze a stable and reliable rotating machinery.

Check in for next week’s blog post – Foundations of Rotor Dynamics: Part 2  and check out AxSTREAM RotorDynamics or SoftInWay’s upcoming webinar on rotordynamics!

Introduction to your Supercritical CO2 Power Cycle

Supercritical carbon dioxide cycles have slowly become more popular in the engineering market for electricity generation from various sources. SCO2 is found to be an ideal working fluid for generating power cycles due to its high efficiency –more than supercritical or superheated steam, which results in lower cost of electricity.

Supercritical carbon dioxide is a fluid state where carbon dioxide is operated above its critical point which causes the compound to behave as both a gas and a liquid simultaneously with the unique ability to flow as a gas though at the same time dissolve materials like a liquid. SCO2 changes density over small difference in temperature or pressure, though stay in the same phase; allowing large amount of energy to be extracted at higher temperatures.
This cycle works in a similar manner to other power generation cycles, and is potentially applicable to wide variety of power generation applications. Hypothetically speaking, any cycle that is running with steam as the working fluid should be able to be upgraded to SCO2 application. In an example for applications using fossil fuel as a main heat source, cycle could be designed as an indirectly-heated non-condensing closed-loop Brayton cycle or directly fired SCO2. In the first event, CO2 is heated non-directly through a heat exchanger. After that, the hot CO2 flow expands in the turbine where the mechanical energy is extracted and any remaining heat is extracted in the recuperator to preheat the CO2 going back to the inlet loop, resulting to high efficiency systems. Where for second arrangement, fossil fuel is directly combusted with oxygen, resulting to steam/CO2 mixture to drive the turbine and generate electricity. The remaining heat in the fluid mixture will be recuperated to preheat the CO2 that is used as the combustion diluent.

There are many benefits that come with SCO2 power conversion technology when compared to other power cycles such as higher efficiency (which correspondent to higher productivity with the same thermal input), environmentally friendly/low greenhouse gas emission, and lower capital cost from reduced size compared to a conventional steam cycle.

Want to optimize your SCO2 cycle? Check out our simulation technology AxCYCLE



Design Challenges of Boiler Feed Pump Turbines in Thermal Power Stations

 The design of a boiler feed pump turbine features some unique characteristics that presents certain challenges in terms of efficiency management, varying operating ranges, and many other features.  In order better understand the accepted designs of Boiler Feed Pump Turbines (BFPTs), it is important to know how the operation of steam turbines used to drive boiler feed pumps can fundamentally improve fossil and nuclear plants.  Much like the design of mechanical drive turbines, feed pump turbines also feature the same thermodynamic objectives as the main turbine and all of the engineering difficulties with optimal blade design, rotor and bearing harmonic conditions, ideal flow path definitions, and so on.  However, some distinctions can make a BFPT design particularly distinct from a regular mechanical drive turbine.  Figure 1 shows a basic heat balance diagram for a plant using a boiler feed pump turbine arrangement.

Figure 1 – Simple Process Diagram for Plant with Boiler Feed Pump Turbine in AxCYCLE®

Inherent in its name, the BFPT must be fully compatible with the boiler feed pump. In other words, the necessary power and speed of the BFPT are determined by the requirements of the pump. In a fully integrated and dynamic system such as this, a large portion of the design requires developing a proper heat balance that will optimize the plant performance. In general, the boiler feed pump turbine uses both steam from the boiler and the main turbine to drive the mechanical shaft connected to the boiler feed pump. This arrangement has proven highly successful in efficiently applying the steam’s thermal energy throughout the plant. In certain arrangements, the BFPT can instead accept steam from cold reheat lines, main unit crossover piping lines, and different extractions from the main turbine. Regardless of the source, one distinction specifically unique to the BFPT is that it must accept steam from two separate sources.

In reference to schematic in Figure 1, the BFPT accepts steam at different pressures from both the boiler and the main turbine.  The low-pressure steam extracted from the main turbine, generally between the high-pressure (HP) turbine and intermediate pressure (IP) turbine sections, will range from 75 psig to 250 psig.  On the other hand, high-pressure steam directed from the boiler can reach pressures as high as 2400 psig, even 3500 psig in supercritical plants.  The ability to utilize two vastly separate steam sources is made possible with the use of two separate inlet designs for the BFPT.  The inlet designs for both the high pressure and the lower pressure sections of the BFPT consist of a series of valves driven by an actuator.  The percentage in which each of these valve sections are open controls the different operating conditions of the plant.  Three main operating points are considered for the feed pump turbine based on solely the lower pressure steam conditions coming from the main turbine.  The conditions with these valves wide open (VWO), 40% of the main unit load (MUL), and the run out point (65% of MUL) all define the operating ranges of this section of the turbine.  The range associated with each of these points allow the engineer to size the correct areas of the LP nozzles.

High-pressure steam from the boiler can be used to start the BFPT without using an auxiliary steam source.  These start up requirements determine the nozzle sizing for the HP steam inlet section.  As seen above, in order to achieve an optimal and efficient design for a BFPT, a number of different intermediate design points must be considered due to the expansive operating range that this particular turbine experiences.  The analysis of different off-design curves becomes crucial in the design of boiler feed pump turbines and is a must for any engineers looking to improve their axial turbine design for boiler feed pump turbines.  To learn more about the full design of process of SoftInWay’s AxSTREAM®, please click here.

Importance and Modelling of Internal Combustion Engine Cooling Systems

In an internal combustion engine, combustion of air and fuel takes place inside the engine cylinder and hot gases are generated with temperature of gases around 2300-2500°C which may result in not only burning of oil film between the moving parts, but also in seizing or welding of the stationery and moving components. This temperature must be reduced such that the engine works at top efficienc,  promoting high volumetric efficiency and ensuring better combustion without compromising the thermal efficiency due to overcooling. Most importantly, the engine needs to function both in the sense of mechanical operation and reliability. In short, cooling is a matter of equalization of internal temperature to prevent local overheating as well as to remove sufficient heat energy to maintain a practical overall working temperature.

It is also important to note that about 20-25% of the total heat generated is used for producing brake power (useful work). The cooling system should be designed to remove 30-35% of total heat and the remaining heat is lost in friction and carried away by exhaust gases.

The design of cooling systems depends on whether the engine is air cooled or liquid cooled. Air cooling is generally used in small engines wherein fins or extended surfaces are provided on the cylinder walls, cylinder head, and so on. Heat generated due to combustion in the engine cylinder will be conducted to the fins and when the air flows over the fins, heat will be dissipated to air. The amount of heat dissipated to air depends upon: Amount of air flowing through the fins, fin surface area and the thermal conductivity of metal used for fins.

In water cooling methods, cooling water jackets are provided around the cylinder, cylinder head, valve seats and so on. When the water is circulated through the jackets, it absorbs heat of combustion. This hot water will then be cooling in the radiator partially by a fan and partially by the flow developed by the forward motion of the vehicle. The cooled water is again recirculated through the water jackets either through a pump or thermos-siphon which is based on the principle of density difference in working fluid.

Figure 1: Cooling water ports in an IC engine cylinder block

Figure 1. shows the cooling water jacket for an IC engine cylinder block. The engine cooling jacket is of complex shape and is influenced by many factors including the shape of the engine block and optimal temperature at which the engine runs. A large cooling jacket would be effective in transporting heat away from the cylinders, but makes the engine bulky and heavier. The cooling water jacket needs to be optimized considering both the cooling effectiveness and engine weight. Hence the flow through the cooling jacket needs to be optimized from the inlet to the outlet covering the lengthwise along the geometry as well as traversing from cylinder block to the head. The optimization is done with the objective of minimizing the fluid pressure loss between inlet and outlet and obtains even distribution of the flow to each cylinder in the engine block and uniform velocities along its flow.

The engine cooling jacket is of complex geometry and performing 3D simulation over this is quite a complex task involving generating the 3D geometry with all the intricate details and preparing the model for performing conjugate heat transfer analysis. As an initial step it is advisable to perform a simple 1D heat and flow network analysis to obtain the heat transfer distribution and data for creating the 3D model using commercial tools such as AxSTREAM NET™.

To know more about how AxSTREAM NET™ can simplify engine cooling system design and analysis, please write to info@softinway.com.

Explaining Geothermal Cycles


Geothermal energy has become more and more popular globally due to its sustainability and economic stand point. Geothermal power plants run on a variety of temperatures and utilize hydrothermal resources (water/steam and heat) from below the earth surface to generate electricity for people’s daily consumption. Resources can come from dry steam or hot water wells.

There are three kinds of Geothermal cycle for power plants: binary cycle, dry steam and flash steam. Binary cycle power plants use the heat transfer from geothermal hot water to secondary fluids with a low boiling point at the lower end of standard geothermal temperature (225 to 360 F). This heat will cause the secondary fluid to bubble and turn into steam in the heat exchanger, which is then used to turn the turbine. Since water and secondary fluids are kept apart in the cycle, air emission is minimized.

Dry steam is the first geothermal power plant to ever exist from a natural rupture of steam, though considerably uncommon since it demands sustainable underground heat sources to work. The steam used as a working fluid will be piped directly from the underground geothermal reservoirs to turn a turbine and generate electricity.

Flash steam is the most common type of geothermal application, using a underground high-pressure hot water reservoir with a minimum temperature of 360 F, converting it to steam as it moves up to the surface from change in pressure. After steam gets separated from water, it drives the turbines to produce electricity. As the steam cools down and condenses to water, fluid then will be injected back to the reservoir to be reused.

Each one of these system designs comes with its own advantages and disadvantages. For example, binary cycle allows low temperature geothermal sources to be used thus can be used in more wide spread applications. This kind of cycle also does not release geothermal fluid into the system, thus the technology is more environmentally friendly. On the other hand flash steam power plant gives you the advantage of sustainability as well as cost effectiveness in the long run, though it’s rather geographically sensitive. Dry steam application is hard to implement due to the rather rare natural resource used to be able to implement such a cycle, though it generates less of a footprint and require simpler technology which results to lower initial cost. The better application is really dependent on the designer’s needs and goals.






Driving Turboexpander Technology

Turboexpanders are used in a number of applications, including floating LNG (liquefied natural gas), LPG (liquefied petroleum gas) / NGL (natural gas liquids), dew point control, and ethylene plants.  Used as a highly efficient system that takes advantage of high pressure, high-temperature flows, the turboexpander both produces cryogenic temperatures and simultaneously converts thermal energy into shaft power.  Essentially, a turboexpander is comprised of a radial inflow expansion turbine and a centrifugal compressor combined as a single unit on a rigid shaft. The process fluid from a plant stream will run through the expansion turbine to both provide low-temperature refrigeration and convert thermal energy to mechanical power as a byproduct.  First, the gas will radially enter the variable inlet nozzles (or guide vanes) of the turbine, which will allow for a localized increase in fluid velocity prior to entering the turbine wheel.  The turbine wheel will accept this high-temperature, high-pressure, accelerated gas and convert it into mechanical energy via shaft rotation. The primary product of a turboexpander manifests at the outflow of this turbine.  After the process gas passes through the turbine wheel, this gas has expanded so dramatically that it produces cryogenic temperatures colder than any other equipment in the plant.

Figure 1- Typical Turboexpander – Expander-Compressor Configuration

The useful mechanical energy converted from this system is generally used to drive a centrifugal compressor positioned on the opposite end of the shaft.  In the case of this expander-compressor setup, the mentioned turboexpander technology avoids the excessive use of fuel consumption seen in other systems, and significantly decreases the CO2 footprint of the overall design.  As well, there are various examples of turboexpanders that use an expander-generator setup, which converts the mechanical energy from the turbine into direct electrical power.  Turboexpanders have come a long way in the last 40 years.  With the advent of magnetic bearings and more advanced sealing systems, turboexpanders have been able to handle shaft speeds in large and small machines of up to 10,000 rpm and 120,000 rpm, respectively.  Moreover, innovations in specific CFD modules for turbomachinery have allowed turboexpander systems to achieve efficiencies upwards of 90%.

Figure 2- Turboexpander – Expander-Generator Configuration

All the liquefaction applications mentioned above use this technology due to the optimal efficiencies it can produce.  An accepted condensation method for gases in the past was a technique called Joule-Thomson cooling.  The Joule-Thomson effect describes the phenomenon that a temperature change can occur in a gas due to a sudden pressure change over a valve.  Although extremely important in the advancement of refrigeration systems, Joule-Thomson cooling exclusively required much higher pressure in order to remove the same amount of energy because no external work is done.  The introduction of mechanical load devices allowed the process gas to distribute its energy more rapidly, in turn allowing more optimal refrigeration conditions.

The flow of gas through an expansion turbine encounters numerous rapid transitions and requires an in depth look using CFD simulations for analysis and further optimization.  To learn more about the design of turboexpander components, please visit http://www.softinway.com/machine-type/centrifugal-compressor/







Using 1D Models to Predict the Thermal Growth and Stresses During The Start up and Shutdown Phase of a Steam Turbine

Steam turbines are not just restricted to conventional or nuclear power plants, they are widely used in combined cycle power plants, concentrated solar thermal plants and also geothermal power plants. The operational requirements of a steam turbine in the combined cycle and CSP’s means that they operate under transient conditions. Even in conventional steam turbines, the market requirements are changing with requirements for faster and more frequent start-up which can result into faster deterioration of the equipment and reduced lifespan. During the startup phase, significant heat exchange takes place between the steam and the structural components that include the valves, rotor and casing. The accuracy of the life prediction is strongly affected and dependent on the accuracy of the transient thermal state prediction [1].

Though the expansion of steam takes place in the nozzles and blades, the influence of the leakage steam during the startup phase is significant with steam expanding through the labyrinths resulting in expansions, condensation, and increased velocities which may even reach supersonic levels. During cold start, the flow is minimal, the temperature of the metal is at room temperature and heat exchange happens between the steam and metal parts resulting in thermal stress.

Every designer is interested in making a prediction that is as accurate as possible. This requires modelling the entire flow path with all the intricate details which means generating a complex 3D model,use of extensive computational resources and so on, which ultimately results in more time and cost. Even if one has the luxury of using a complex 3D model with all the intricate details, the question is how to get the appropriate boundary conditions to be applied for the 3D simulation and how to reduce the number of iterations required between the flow analysis and structural analysis. The flow parameters and the area between sealing fins in the labyrinth (refer Fig.1) is varied, not only in each stage, but also within each component which means the heat transfer coefficient being applied to each of these locations is also different.

   Figure 1: Sealing fins in a turbine stage


As seen in Fig. 2, the entire flow through the seals and cavities can be modelled as a 1D model considering both the flow and heat transfer between the fluid and metal. During design phase of the flow path, a streamline analysis tool such as AxSTREAM® can give the leakage flow parameters which can be further detailed using AxSTREAM® NET, the 1D flow and heat transfer module. The results from the 1D module, which gives the thermodynamic parameters and heat transfer coefficients at different zones in the seal, can be used to apply the boundary conditions for performing thermos-structural analysis of the steam turbine.

                    Figure 2: 1D model of the flow and heat transfer in sealing fins of a turbine stage


To learn more about how the AxSTREAM® platform can be used for designing steam turbines, from concept to optimizing operating curves, please contact info@SoftInWay.com



Understanding the Characteristics of Varying Centrifugal Blower Designs

Many people speculate about the confusion on what is considered a compressor, a blower, or simply a fan.  In essence, each of these turbo-machines achieve a pressure rise by adding velocity to a continuous flow of fluid.  The distinctions between fans, blowers, and compressors are quite simply defined by one parameter, the specific pressure ratio.  Each machine type, however, utilizes a number of different design techniques specific to lower and higher-pressure applications.  As per the American Society of Mechanical Engineers (ASME), the specific pressure is defined as the ratio of the discharge pressure over the suction pressure (or inlet pressure).  The table shown below defines the range at which fans, blowers, and compressors are categorized.

Similarities between the design of fans and blowers occur near the lower end of a blower’s range.  As well, many design parallels exist between high-pressure blowers and compressors.  For the article, we will be investigating the different design characteristics of centrifugal blowers. Blower selection depends on a number of factors including operating range, efficiency, space limitations, and material handled.   Figure 1 shows a number of different impeller blade designs that are available for centrifugal blowers.

Figure 1 – Impeller Blading Arrangements for Centrifugal Blowers

Each of these blading arrangements have unique operating ranges limited by overload and stall conditions.  Forward curve and radial blowers exhibit a performance curve for horsepower needed that increases with the volume flow rate.  This means that a motor can be overloaded if discrepancies occur which bring the system to flow rates higher than at the operating point.  On the other hand, any backward arrangement blading exhibits a horsepower curve that increases to a maximum as airflow increases and then drops off again.  This allows the engineer to specify a motor to accommodate the peak horsepower, which in turn ensures that the system cannot overload and is therefore considered “non-overloading”.  The stable range of a blower is defined as the condition under which enough air flows through the fan wheel to fill the spaces in between the blades.  Below this range, the instability of the machine will cause one or several section of blades to stall.

In general, forward curved blading arrangements (or sirocco blowers) are better suited for high volume with lower pressure applications.  These blowers and fans operate at relatively low speeds and pressures, which permit lightweight and cost-effective construction.  Restricted stability ranges, overloading at higher flow rates, and low static efficiencies all limit the capabilities of the forward curved centrifugal blower.  Radial centrifugal blowers have the same problem with overloading at high flow rates; however, they are quite beneficial for moving air in dirty environments due to their flat geometry that prohibits dust or sticky materials from rapidly accumulating.  Composed of 6 to 12 rugged blades extending radially from the hub, these types of blowers run at medium speeds and deliver low air volumes at medium to high pressure.

Figure 2 – Centrifugal Blower Design with Airfoil Impeller Arrangement in AxSTREAM™

Perhaps the most important centrifugal fan impeller orientation, the backward orientation, comes in three standard shapes: backward inclined, backward curved, and backward inclined aerofoil (or airfoil). All of these designs exhibit most of the same characteristics, with some discrepancies in their efficiencies.  Generally, the “flat-bladed” back inclined design achieves an efficiency of about 82%, while the backward curved and airfoil designs near 88% and 90%, respectively.  In addition to high efficiencies, backward oriented blade designs allow for the highest operating speeds of all centrifugal blowers and are considered non-overloading.  The only drawbacks to these types of blowers would be the high cost of manufacturing as well as the inability to handle flows with high particulates due to the close running clearances and complex geometries.

If you would like to learn more about the design of centrifugal fans, blowers, and compressors, please click here (http://www.softinway.com/machine-type/centrifugal-compressor/).





Gaining Turbomachinery Insight Using a Fluid Structure Interaction Approach

Existing research studies for the corresponding flow-induced vibration analysis of centrifugal pumps are mainly carried out without considering the interaction between fluid and structure. The ignorance of fluid structure interaction (FSI) means that the energy transfer between fluid and structure is neglected. To some extent, the accuracy and reliability of unsteady flow and rotor deflection analysis should be affected by this interaction mechanism.

In recent years, more and more applications of FSI are found in the reliability research of turbomachinery. Most of them are about turbines, and a few of them address pumps. Kato [1] predicted the noise from a multi-stage centrifugal pump using one-way coupling method. This practical approach treats the fluid physics and the solid physics consecutively.

Figure 1: Multistage centrifugal pump [1].
In the CFD computations of the internal flows, Kato could successfully predict the pressure fluctuations despite turbulent boundary layer in the impeller passages was not resolved. The computed pressure fluctuations on the internal surface agreed well with the measured ones not only at the blade passing frequencies, (BPF) but also on the base level. By visualizing the distributions of the pressure fluctuations at the BPFs, it was found that the fluctuation was especially high at the second harmonics of the BPF. This was consistent with the vibration velocity measured on the outer surface. On the other hand, he overpredicted the total head by about 10%. This is because turbulent boundary layer in the impeller passage was not resolved, and therefore, the blockage effect was not taken into account appropriately at this stage of the research.

Vibration of the structure portion was then calculated by a dynamical structural analysis with the calculated pressure fluctuations on the internal surface as input data. It was clearly shown that the dominant vibrations of the pump originate from the rotor-stator interaction. The trivial vibrations were damped off over time. The vibration levels of the BPF on the outer surface of the pump structure agreed reasonably well with the  measured ones. The computations revealed the feasibility of the fluid-structure coupled simulation for flow-induced noise generated in turbomachinery.

Another example of fluid-structure interaction was presented by Pei et. Al [2] when an axial-flow pump device with a two-way passage was studied. A coupled solution of the flow field and structural response of the impeller was established using a two-way coupling method to study the distribution of stress and deformation in the impeller and quantitatively analyze that on the blade along the wireframe paths had different flow rates. This studied showed that the maximum equivalent stress and maximum total deformation in the impeller are greatly influenced by flow rate, and its values drops with an increasing flow rate and a decreasing head. In addition, the total deformation in the impeller is greater near the blade rim, where the maximum value can be found. The equivalent stress is greater near the blade hub, where the maximum value can be obtained.

The above studies are the best proof that by using the right methods, tools and expertise you can get an insight for any kind of turbomachinery. Try AxSTREAM using the CFD and FEA integrated modules to design your machine and understand the fundamentals of its operation in depth.


[1] Prediction of the Noise From a Multi-Stage Centrifugal Pump, Chisachi Kato, Shinobu Yoshimura, Yoshinobu Yamade, Yu Yan Jiang, Hong Wang, Ryuta Imai, Hiroyuki Katsura, Tetsuya Yoshida and Yashushi Takano , ASME 2005 Fluids Engineering Division Summer Meeting, Volume 1: Symposia, Parts A and B, Houston, Texas, USA, June 19–23, 2005

[2] Fluid–structure coupling analysis of deformation and stress in impeller of an axial-flow pump with two-way passage, Ji Pei, Fan Meng, Yanjun Li, Shouqi Yuan, Jia Chen, National Research Center of Pumps, Jiangsu University, Zhenjiang, China