Centrifugal compressor

thumb|upright= 1.35|Centrifugal impeller, shown alone

thumb|upright= 1.35| Centrifugal compressor shown (in blue) as part of a [[turbocharger]]

thumb|upright= 1.35|Centrifugal compressor shown (in blue) as second stage of an axi-centrifugal [[jet-engine]]

Centrifugal compressors, sometimes called impeller compressors or radial compressors, are a sub-class of dynamic, axisymmetric, work-absorbing turbomachinery.

They achieve pressure rise by adding energy to the continuous flow of fluid through the rotor/impeller. The equation in the next section shows this specific energy input. A substantial portion of this energy is kinetic, which is converted to increased potential energy/static pressure by slowing the flow through a diffuser. The static pressure rise in the impeller may roughly equal the rise in the diffuser.

Components of a simple centrifugal compressor

thumb|upright= 1.35|Figure 1.1Two-stage turboshaft, first-stage flowpath, annular inlet, guide vanes, open impeller, vaned diffuser, vaneless return bend

A simple centrifugal compressor stage has four components (listed in order of throughflow): inlet, impeller/rotor, diffuser, and collector.

Derived from the general Euler equations is Euler's pump and turbine equation, which plays an important role in understanding impeller performance. This equation can be written in the form:

Subscript 1 is the impeller leading edge (inlet), station 1
Subscript 2 is the impeller trailing edge (discharge), station 2
is the energy added to the fluid
is the acceleration due to gravity
is the impeller's circumferential velocity, units velocity
is the velocity of flow relative to the impeller, units velocity
is the absolute velocity of flow relative to stationary, units velocity

Impeller inlet meridional triangles.PNG|Figure 1.2.2Inlet velocity triangles for centrifugal compressor impeller

Impeller exit meridional trianges.PNG|Figure 1.2.3Exit velocity triangles for centrifugal compressor impeller

</gallery>

Diffuser

thumb|upright= 1.35|Figure 1.3NASA CC3 impeller and wedge diffuser

The next component downstream of the impeller within a simple centrifugal compressor may be the diffuser.

Bernoulli's principle is important in understanding how diffusers perform. In engineering situations assuming adiabatic flow, this equation can be written in the form:

where:

Subscript is the inlet of the diffuser, station 4
Subscript is the discharge of the diffuser, station 5
(see inlet above.)

Historical contributions, the pioneers

Over the past 100 years, applied scientists including Stodola (1903, 1927–1945), Pfleiderer (1952), Hawthorne (1964), Shepherd (1956), and Japikse (many texts including citations), Improvements in centrifugal compressors have not been achieved through large discoveries. Rather, improvements have been achieved through understanding and applying incremental pieces of knowledge discovered by many individuals.

Aerodynamic-thermodynamic domain

thumb|upright= 1.35|Figure 2.1Aero-thermo domain of turbomachinery

Figure 2.1 (shown right) represents the aero-thermo domain of turbomachinery. The horizontal axis represents the energy equation derivable from the first law of thermodynamics. Kernelien Le Demour, Daniel Gabriel Fahrenheit, John Smeaton, Dr. A. C. E. Rateau, John Barber, Alexander Sablukov, Sir Charles Algernon Parsons, Ægidius Elling, Sanford Alexander Moss, Willis Carrier, Adolf Busemann, Hermann Schlichting, Frank Whittle, and Hans von Ohain.

Partial timeline of historical contributions

{| class="wikitable"

|+ Table 2.1

| <1689

| Early turbomachines

| Pumps, blowers, fans

| 1689

| Denis Papin

| Origin of the centrifugal compressor

| 1754

| Leonhard Euler

| Euler's pump-and-turbine equation

| 1791

| John Barber

| First gas-turbine patent

| 1899

| A. C. E. Rateau

| First practical centrifugal compressor

| 1927

| Aurel Boleslav Stodola

| Formalized "slip factor"

| 1928

| Adolf Busemann

| Derived "slip factor"

| 1937

| Frank Whittle and Hans von Ohain, independently

| First gas turbine using a centrifugal compressor

| >1970

| Modern turbomachines

| 3D-CFD, rocket turbo-pumps, heart assist pumps, turbocharged fuel cells

Turbomachinery similarities

Centrifugal compressors are similar in many ways to other turbomachinery and are compared and contrasted as follows:

Similarities to axial compressor

thumb|upright= 1.35|Cutaway showing an axi-centrifugal compressor gas turbine

Centrifugal compressors are similar to axial compressors in that they are rotating airfoil-based compressors. Both are shown in the adjacent photograph of an engine with 5 stages of axial compressors and one stage of a centrifugal compressor. in the Pratt & Whitney Canada PW200 series of helicopter engines) than does an axial stage. The 1940s-era German Heinkel HeS 011 experimental engine was the first aviation turbojet to have a compressor stage with radial flow-turning part-way between none for an axial and 90 degrees for a centrifugal. It is known as a mixed/diagonal-flow compressor. A diagonal stage is used in the Pratt & Whitney Canada PW600 series of small turbofans.

Centrifugal fan

thumb|upright=1.3|A low-speed, low-pressure centrifugal compressor or [[centrifugal fan, with upward discharging cone used to diffuse the air velocity]]

Centrifugal compressors are also similar to centrifugal fans of the style shown in the neighboring figure, as they both increase the energy of the flow through the increasing radius. when the working fluid is air or nitrogen. In contrast, fans or blowers are often considered to have density increases of less than five percent and peak relative fluid velocities below Mach 0.3.

Squirrel-cage fan

thumb|upright=1.3|A low-speed, low-pressure blower used for HVAC ventilation

Squirrel-cage fans are primarily used for ventilation. The flow field within this type of fan has internal recirculations. In comparison, a centrifugal fan is uniform circumferentially.

Centrifugal pump

Centrifugal compressors are also similar to centrifugal pumps

American Petroleum Institute: API STD 617 8TH ED (E1), API STD 672 5TH ED (2019).
American Society of Heating, Refrigeration, and Airconditioning Engineers: Handbook Fundamentals.
Society of Automotive Engineers
Compressed Air and Gas Institute
International Organization for Standardization: ISO 10439, ISO 10442, ISO 18740, ISO 6368, ISO 5389

Applications

Below is a partial list of centrifugal-compressor applications, each with a brief description of some of the general characteristics possessed by those compressors. To start this list, two of the most well-known centrifugal compressor applications are listed: gas turbines and turbochargers. Ref. Figures 4.1–4.2 In their simple form, modern gas turbines operate on the Brayton cycle. (ref Figure 5.1) Either or both axial and centrifugal compressors are used to provide compression. The types of gas turbines that most often include centrifugal compressors include small aircraft engines (i.e. turboshafts, turboprops, and turbofans), auxiliary power units, and micro-turbines. The industry standards applied to all centrifugal compressors used in aircraft applications are set by the relevant civilian and military certification authorities to achieve the safety and durability required in service. Centrifugal impellers used in gas turbines are commonly made from titanium alloy forgings. Their flow-path blades are commonly flank-milled or point-milled on 5-axis milling machines. When running clearances have to be as small as possible without the impeller rubbing its shroud, the impeller is first drawn with its high-temperature, high-speed deflected shape and then drawn in its equivalent cold, static shape for manufacturing. This is necessary because the impeller deflections at the most severe running condition can be 100 times larger than the required hot running clearance between the impeller and its shroud.

In automotive engine and diesel engine turbochargers and superchargers. Ref. Figure 1.1 Centrifugal compressors used in conjunction with reciprocating internal combustion engines are known as turbochargers if driven by the engine's exhaust gas and turbo-superchargers if mechanically driven by the engine. Standards set by the industry for turbochargers may have been established by SAE. Centrifugal compressors for such uses are often one-shaft, multi-stage, and driven by large steam or gas turbines. Their casings are termed horizontally split if the rotor is lowered into the bottom half during assembly or barrel if it has no lengthwise split-line with the rotor being slid in. Standards set by the industry (ANSI/API, ASME) for these compressors result in thick casings to achieve a required level of safety. The impellers are often of the covered style, which makes them look much like pump impellers. This type of compressor is also often termed API-style. The power needed to drive these compressors is usually in the thousands of horsepower. Use of real-gas properties is needed to properly design, test and analyze their performance.

Air-conditioning and refrigeration and HVAC: Centrifugal compressors often supply the compression in water-chiller cycles. Because of the wide variety of vapor compression cycles (thermodynamic cycle, thermodynamics) and the wide variety of working fluids (refrigerants), centrifugal compressors are used in a variety of sizes and configurations. Use of real-gas properties is needed to properly design, test, and analyze the performance of these machines. Standards set by the industry for these compressors include ASHRAE, ASME, & API.

In industry and manufacturing to supply compressed air for all types of pneumatic tools. Centrifugal compressors for such uses are often multistage, using inter-cooling to control air temperature. Standards set by the industry for these compressors include ASME and government regulations that emphasize safety. Ideal-gas relationships are often used to properly design, test, and analyze the performance of these machines when the working gas is air or nitrogen. Other gases require real-gas properties.

In oil field re-injection of high-pressure natural gas to improve oil recovery. Figure 5.1 includes example plots of pressure-specific volume and temperature-entropy. These types of plots are fundamental to understanding centrifugal compressor performance at one operating point. The two plots show that the pressure rises between the compressor inlet (station 1) and compressor exit (station 2). At the same time, the specific volume decreases while the density increases. The temperature-entropy plot shows that the temperature increases with increasing entropy (loss). Assuming dry air, and the ideal-gas equation of state and an isentropic process, there is enough information to define the pressure ratio and efficiency for this one point. The compressor map is required to understand the compressor performance over its complete operating range.

Figure 5.2, a centrifugal-compressor performance map (either test or estimated), shows the flow, pressure ratio for each of 4 speed-lines (total of 23 data points). Also included are constant-efficiency contours. Centrifugal-compressor performance presented in this form provides enough information to match the hardware represented by the map to a simple set of end-user requirements.

Compared to estimating performance, which is very cost effective (thus useful in design), testing, while costly, is still the most precise method. Further, testing centrifugal compressor performance is very complex. Professional societies such as ASME (i.e. PTC–10, Fluid Meters Handbook, PTC-19.x), ASHRAE (ASHRAE Handbook) and API (ANSI/API 617–2002, 672–2007) It is standard in these cases that the equivalent temperature, equivalent pressure, and gas is specified explicitly or implied at a standard condition.

Volume flow per unit time

In contrast, all volume-flow specifications require the additional specification of density. Bernoulli's principle is of great value in understanding this problem. Confusion arises through either inaccuracies or misuse of pressure, temperature, and gas constants.

Also as is standard practice, Figure 5.2 has a vertical axis labeled with a pressure parameter. There is a variety of pressure measurement units. They all fit one of two categories:

A change in pressure, ie increase from inlet to exit (measured with a manometer)
A discharge pressure

The pressure rise may alternatively be specified as a ratio that has no units:

A pressure ratio (exit/inlet)

Other features common to performance maps are constant-speed lines, constant-efficiency islands, and design or guarantee points.

Constant-speed lines

The two most common methods for producing a map for a centrifugal compressor are at constant shaft speed or with a constant throttle setting. If the speed is held constant, then test points are taken along a constant-speed line by changing throttle positions. In contrast, if a throttle valve is held constant, then test points are established by changing speed and repeated with different throttle positions (common gas turbine practice). The map shown in Figure 5.2 illustrates the most common method; lines of constant speed. In this case, we see data points connected via straight lines at speeds of 50%, 71%, 87%, and 100% RPM. The first three speed-lines have 6 points each, while the highest speed line has five.

Constant-efficiency islands

The next feature to be discussed is the oval-shaped curves representing islands of constant efficiency. In this figure, we see 11 contours ranging from 56% efficiency (decimal 0.56) to 76% efficiency (decimal 0.76). General standard practice is to interpret these efficiencies as isentropic rather than polytropic. The inclusion of efficiency islands effectively generates a 3-dimensional topology for this 2-dimensional map. With inlet density specified, it provides a further ability to calculate aerodynamic power. Lines of constant power could just as easily be substituted.

Design or guarantee points

Regarding gas turbine operation and performance, there may be a series of guaranteed points established for the gas turbine's centrifugal compressor. These requirements are of secondary importance to the overall gas turbine performance as a whole. For this reason, it is only necessary to summarize that in the ideal case, the lowest specific fuel consumption would occur when the centrifugal compressor's peak efficiency curve coincides with the gas turbine's required operation line.

In contrast to gas turbines, most other applications (including industrial) need to meet a less stringent set of performance requirements. Historically, centrifugal compressors applied to industrial applications were needed to achieve performance at a specific flow and pressure. Modern industrial compressors are often needed to achieve specific performance goals across a range of flows and pressures, thus taking a significant step toward the sophistication seen in gas turbine applications.

If the compressor represented in Figure 5.2 is used in a simple application, then any point (pressure and flow) within the 76% efficiency would provide very acceptable performance. An end user would be very happy with the performance requirements of 2.0 pressure ratio at 0.21 kg/s.

Surge

Surge is a low-flow phenomenon where the impeller cannot add enough energy to overcome the system resistance or backpressure. At low-flow-rate operation, the pressure ratio over the impeller is high, as is backpressure. Under critical conditions, the flow will reverse back over the tips of the rotor blades towards the impeller eye (inlet). This stalling flow reversal may go unnoticed if the fraction of mass flow or energy is too low. When large enough, rapid flow reversal occurs (i.e., surge). The reversed flow exiting the impeller inlet exhibits a strong rotational component, which affects lower-radius flow angles (closer to the impeller hub) at the leading edge of the blades. The deterioration of the flow angles causes the impeller to be inefficient. A full flow reversal can occur. (Therefore, surge is sometimes referred to as axisymmetric stall.) When reversed flow reduces to a low-enough level, the impeller recovers and regains stability for a short moment, at which point the stage may surge again. These cyclic events cause large vibrations, increase temperature, and change rapidly the axial thrust. These occurrences can damage the rotor seals, rotor bearings, the compressor driver, and cycle operation. Most turbomachines are designed to easily withstand occasional surging. However, if the machine is forced to surge repeatedly for a long period of time, or if it is poorly designed, then repeated surges can result in a catastrophic failure. Of particular interest is that while turbomachines may be very durable, their physical system can be far less robust.

Surge line

thumb|upright=1.3|Figure 6.2.1: Stall formation

The surge-line shown in Figure 5.2 is the curve that passes through the lowest flow points of each of the four speed-lines. As a test map, these points would be the lowest flow points possible to record a stable reading within the test facility/rig. In many industrial applications, it may be necessary to increase the stall line due to the system backpressure. For example, at 100% RPM stalling flow might increase from approximately 0.170 kg/s to 0.215 kg/s because of the positive slope of the pressure ratio curve.

As stated earlier, the reason for this is that the high-speed line in Figure 5.2 exhibits a stalling characteristic or positive slope within that range of flows. When placed in a different system, those lower flows might not be achievable because of interaction with that system. System resistance or adverse pressure is proven mathematically to be the critical contributor to compressor surge.

Maximum flow line versus choke

Choke occurs under one of two conditions. Typically for high-speed equipment, as flow increases, the velocity of the flow can approach sonic speed somewhere within the compressor stage. This location may occur at the impeller inlet "throat" or at the vaned diffuser inlet "throat". In contrast, for lower-speed equipment, as flows increase, losses increase such that the pressure ratio eventually drops to 1:1. In this case, the occurrence of choke is unlikely.

The speed-lines of gas-turbine centrifugal compressors typically exhibit choke. This is a situation where the pressure ratio of a speed line drops rapidly (vertically) with little or no change in flow. In most cases the reason for this is that close-to-Mach-1 velocities have been reached somewhere within the impeller or diffuser, generating a rapid increase in losses. Higher-pressure-ratio turbocharger centrifugal compressors exhibit this same phenomenon. Real choke phenomena are a function of compressibility as measured by the local Mach number within an area restriction within the centrifugal pressure stage.

The maximum flow line, shown in Figure 5.2, is the curve that passes through the highest-flow points of each speed line. Upon inspection, it may be noticed that each of these points has been taken near 56% efficiency. Selecting a low efficiency (<60%) is the most common practice used to terminate compressor performance maps at high flows. Another factor that is used to establish the maximum flow line is a pressure ratio near or equal to 1. The 50% speed line may be considered an example of this.

The shape of Figure 5.2's speed-lines provides a good example of why it is inappropriate to use the term choke in association with a maximum flow of all centrifugal compressor speed-lines. In summary, most industrial and commercial centrifugal compressors are selected or designed to operate at or near their highest efficiencies and to avoid operation at low efficiencies. For this reason, there is seldom a reason to illustrate centrifugal compressor performance below 60% efficiency.

Many industrial and commercial multistage compressor performance maps exhibits this same vertical characteristic for a different reason related to what is known as stage stacking.

Other operating limits

;Minimum operating speed: The minimum speed for acceptable operation; below this value, the compressor may be controlled to stop or go into an idle condition.

;Maximum allowable speed: The maximum operating speed for the compressor. Beyond this value, stresses may rise above prescribed limits, and rotor vibrations may increase rapidly. At speeds above this level, the equipment will likely become very dangerous and be controlled to lower speeds.

Dimensional analysis

To weigh the advantages between centrifugal compressors, it is important to compare 8 parameters classic to turbomachinery: pressure rise (p), flow (Q), angular speed (N), power (P), density (ρ), diameter (D), viscosity (μ), and elasticity (e). This creates a practical problem when trying to experimentally determine the effect of any one parameter. This is because it is nearly impossible to change one of these parameters independently.

The Buckingham π theorem can help solve this problem by generating 5 dimensionless forms of these parameters. These parameters provide the foundation for "similitude" and the "affinity laws" in turbomachinery. They provide for the creation of additional relationships (being dimensionless) found valuable in the characterization of performance.

For the example below, head will be substituted for pressure, and sonic velocity will be substituted for elasticity.

Buckingham Π theorem

The three independent dimensions used in this procedure for turbomachinery are:

<math>M</math> mass (force is an alternative)
<math>L</math> length
<math>T</math> time

According to the theorem, each of the eight main parameters are equated to its independent dimensions as follows:

{| class="wikitable"

| Flow

| <math> Q = </math>

| <math> \frac{L^3}{T}</math>

| ex. = m/s

| Head

| <math> H = </math>

| <math> \frac{M L}{T^2}</math>

| ex. = kg·m/s

| Speed

| <math> U = </math>

| <math> \frac{L}{T}</math>

| ex. = m/s

| Power

| <math> P = </math>

| <math> \frac{M L^2}{T^3}</math>

| ex. = kg·m/s

| Density

| <math> \rho = </math>

| <math> \frac{M}{L^3}</math>

| ex. = kg/m

| Viscosity

| <math> \mu = </math>

| <math> \frac{M}{L T}</math>

| ex. = kg/m·s

| Diameter

| <math> D = </math>

| <math> L</math>

| ex. = m

| Speed of sound

| <math> a = </math>

| <math> \frac{L}{T}</math>

| ex. = m/s

Classic turbomachinery similitude

Completing the task of following the formal procedure results in generating this classic set of five dimensionless parameters for turbomachinery.

{| class="wikitable"

| 1

| Corrected mass flow coefficient

| <math> \Pi_1 \Pi_4 = </math>

| <math> \frac{m}{p D^2} \left(\frac{Rt}{k}\right)^{0.5}</math>

| 2

| Alternate#1 equivalent Mach form

| <math> \frac{m_1}{\rho_1 a_1 {D_1}^2 \cos(\alpha_1)} = </math>

| <math> \frac{m_2}{\rho_2 a_2 {D_2}^2 \cos(\alpha_2)}</math>

| 3

| Alternate#2 simplified dimensional form

| <math> \frac{m_1 {t_1}^{0.5{p_1} = </math>

| <math> \frac{m_2 {t_2}^{0.5{p_2}</math>

| 4

| Specific speed coefficient

| <math> \frac = </math>

| <math> \frac{N Q^{0.5{(g H)^{0.75</math>

| 5

| Specific diameter coefficient

| <math> \frac = </math>

| <math> \frac{D (g H)^{0.25{Q^{0.5</math>

It may be found interesting that the specific speed coefficient may be used in place of speed to define the y-axis of Figure 1.2, while at the same time, the specific diameter coefficient may be in place of diameter to define the z-axis.

Affinity laws

The following affinity laws are derived from the five Π-parameters shown above. They provide a simple basis for scaling turbomachinery from one application to the next.

{| class="wikitable"

| From flow coefficient

| <math> \frac{Q_1}{Q_2} = </math>

| <math> \frac{N_1}{N_2} = </math>

| <math> \left(\frac{D_1}{D_2}\right)^3</math>

| From head coefficient

| <math> \frac{H_1}{H_2} = </math>

| <math> \left(\frac{N_1}{N_2}\right)^2 = </math>

| <math> \left(\frac{D_1}{D_2}\right)^2</math>

| From power coefficient

| <math> \frac{P_1}{P_2} = </math>

| <math> \left(\frac{N_1}{N_2}\right)^3 =</math>

| <math> \left(\frac{D_1}{D_2}\right)^5</math>

Aero-thermodynamic fundamentals

The following equations outline a fully three-dimensional mathematical problem that is very difficult to solve even with simplifying assumptions. Until recently, limitations in computational power forced these equations to be simplified to an inviscid two-dimensional problem with pseudo losses. Before the advent of computers, these equations were almost always simplified to a one-dimensional problem. Solving this one-dimensional problem is still valuable today and is often termed mean-line analysis. Even with all of this simplification, it still requires large textbooks to outline and large computer programs to solve practically.

Conservation of mass

Also termed continuity, this fundamental equation written in general form is as follows:

:<math>\frac{\partial \rho}{\partial t} + \nabla \cdot (\rho \mathbf{v}) = 0</math>

Conservation of momentum

Also termed the Navier–Stokes equations, this fundamental is derivable from Newton's second law when applied to fluid motion. Written in compressible form for a Newtonian fluid, this equation may be written as follows:

:<math>

\rho\left(\frac{\partial\mathbf{v{\partial t} + \mathbf{v} \cdot \nabla\mathbf{v}\right) =

-\nabla p + \mu\nabla^2\mathbf{v} + \left( \frac{1}{3} \mu + \mu^v\right) \nabla\left(\nabla \cdot \mathbf{v} \right) + \mathbf{f}

</math>

Conservation of energy

The first law of thermodynamics is the statement of the conservation of energy. Under specific conditions, the operation of a centrifugal compressor is considered a reversible process. For a reversible process, the total amount of heat added to a system can be expressed as <math>\delta Q=TdS</math>, where <math>T</math> is temperature and <math>S</math> is entropy. Therefore, for a reversible process:

:<math>dU=TdS-pdV.\,</math>

Since , , and are thermodynamic functions of state, the above relation holds also for non-reversible changes. The above equation is known as the fundamental thermodynamic relation.

Equation of state

The classical ideal gas law may be written:

:<math>{\ pV = nRT}.</math>

The ideal gas law may also be expressed as

:<math>{\ p = \rho (\gamma-1)U},</math>

where <math>\rho</math> is the density, <math>\gamma = C_p/C_v</math> is the adiabatic index (ratio of specific heats), <math>U = C_vT</math> is the internal energy per unit mass (the "specific internal energy"), <math>C_v</math> is the specific heat at constant volume, and <math>C_p</math> is the specific heat at constant pressure.

With regard to the equation of state, it is important to remember that while air and nitrogen properties (near standard atmospheric conditions) are easily and accurately estimated by this simple relationship, there are many centrifugal compressor applications where the ideal relationship is inadequate. For example, centrifugal compressors used for large air-conditioning systems (water chillers) use a refrigerant as a working gas that cannot be modeled as an ideal gas. Another example is centrifugal compressors designed and built for the petroleum industry. Most of the hydrocarbon gases, such as methane and ethylene, are best modeled with a real-gas equation of state rather than the ideal gas law.

Pros and cons

;Pros

Centrifugal compressors offer the advantages of simplicity of manufacturing and relatively low cost. This is due to requiring fewer stages to achieve the same pressure rise.
Centrifugal compressors are used throughout industry because they have fewer rubbing parts, are relatively energy-efficient, and give higher and non-oscillating constant airflow than a similarly-sized reciprocating compressor or any other positive-displacement pump.
Centrifugal compressors are mostly used as turbochargers and in small gas-turbine engines like in an auxiliary power unit and as main engine for smaller aircraft like helicopters. A significant reason for this is that with current technology, the equivalent airflow axial compressor will be less efficient due primarily to a combination of rotor and variable stator tip-clearance losses.

;Cons

Their main drawback is that they cannot achieve the high compression ratio of reciprocating compressors without multiple stages. There are few one-stage centrifugal compressors capable of pressure ratios over 10:1, due to stress considerations which severely limit the compressor's safety, durability, and life expectancy.
Centrifugal compressors are impractical, compared to axial compressors, for use in large gas turbines and turbojet engines propelling large aircraft, due to the resulting weight and stress, and to the frontal area presented by the large diameter of the radial diffuser.

Structural mechanics, manufacture, and design compromise

Ideally, centrifugal compressor impellers have thin and strong airfoil blades, each mounted on a light rotor. This material would be easy and cheap to machine or cast. Additionally, it would generate no operating noise, and have a long life while operating in any environment.

From the very start of the aero-thermodynamic design process, the aerodynamic considerations and optimizations [29,30] are critical to have a successful design. During the design phase, the centrifugal impeller's material and manufacturing method must be accounted for, whether it be plastic for a vacuum cleaner blower, aluminum alloy for a turbocharger, steel alloy for an air compressor, or titanium alloy for a gas turbine. It is a combination of the centrifugal-compressor impeller shape, its operating environment, its material, and its manufacturing method that determines the impeller's structural integrity.

References

External links

MIT Gas Turbine Laboratory
(1948), First Marine Gas Turbine in Service. Journal of the American Society for Naval Engineers, 60: 66–86.
A history of Chrysler turbine cars
To find API codes, standards & publications
To find ASME codes, standards & publications
To find ASHRAE codes, standards & publications
Glenn Research Center at NASA
Hydrodynamics of Pumps, by Christopher Earls Brennen
Ctrend website to calculate the head of centrifugal compressor online

pt:Compressor#Compressores Dinâmicos

ru:Лопастной компрессор#Центробежный компрессор

Components of a simple centrifugal compressor

Diffuser

Historical contributions, the pioneers

Aerodynamic-thermodynamic domain

Partial timeline of historical contributions

Turbomachinery similarities

Similarities to axial compressor

Centrifugal fan

Squirrel-cage fan

Centrifugal pump

Applications

Volume flow per unit time

Constant-speed lines

Constant-efficiency islands

Design or guarantee points

Surge

Surge line

Maximum flow line versus choke

Other operating limits

Dimensional analysis

Buckingham Π theorem

Classic turbomachinery similitude

Affinity laws

Aero-thermodynamic fundamentals

Conservation of mass

Conservation of momentum

Conservation of energy

Equation of state

Pros and cons

Structural mechanics, manufacture, and design compromise

See also

References

External links