FR2975122A1 - ROTOR WITH ASYMMETRIC SPACING OF BLADES - Google Patents

Abstract

A turbine apparatus that includes a rotor (30) with a hub section (32) defined about an axis of rotation (A) and a plurality of blades attached to the hub section (32). The plurality of blades comprises a first group having a first angular spacing (S1) in a first circumferential sector (38A) of the rotor (30), and a second group having a second angular spacing (S2) in a second circumferential sector (38B). of the rotor (30). The first angular spacing (S1) is different from the second angular spacing (S2), and the rotor blades are asymmetrical about the axis of rotation (A).

Description

The present invention generally relates to turbine and compressor systems, and particularly to rotor noise reduction. In particular, the invention relates to noise reduction for rotors, compressor wheels and other turbomachine components, including compressor and wheel rotors for auxiliary power units or APUs ("Auxiliary Power Units"). ). Turbine engines are used in a wide range of applications including electric power generation, aviation, and industrial heating and cooling. The turbine core is constructed around a fluidic series compressor section with a combustion chamber and a turbine section, with an upstream inlet and a downstream exhaust. The compressor section compresses air from the inlet, which is mixed with a fuel in the combustion chamber and ignited to generate a hot combustion gas. The turbine section extracts energy from the expanding combustion gas, which is then exhausted through the exhaust. Some of the energy is used to drive the compressor section, and the excess is distributed as a rotational or propulsive thrust, or a combination thereof. The compressor and turbine sections each include a number of rotor blade and stator blade profiles, which are arranged in a series of alternating stages of blades and vanes. In large-scale motors, two, three or more sections may be arranged coaxially in bodies of high pressure, low pressure and intermediate pressure, which can rotate at different speeds, and in different directions. In industrial floor-mounted gas turbines, the drive or output shaft is coupled with an external load such as an electric generator, or with a pump, blower, or other rotating apparatus. In aviation applications, the low body is coupled with a propeller (turboprop) or a propulsion fan (turbo-fan machines), or with a helicopter rotor or rotary wing (turboshaft engines). Depending on the configuration, the coupling may include a gearbox for speed control independent of the fan or output shaft, relative to the low body. The auxiliary power units incorporate smaller scale (typically one-body) gas turbine engines, and are used to generate electrical power and operate various auxiliary systems and accessories. In aviation applications, APUs generate electrical power and provide cabin air when the aircraft is on the ground, and provide compressed air to the sampling system for main engine starting. Depending on the configuration, the APUs can also be used as in-flight power sources for air conditioning and other environmental control systems, and provide independent or emergency 3 backup power for hydraulic, pneumatic, aeronautical electronic elements and piloting. In each of these applications, the compressor and turbine rotors operate at high speeds. When parts of the blades reach transonic and supersonic velocities, they generate shock waves at the frequency due to blade wakes ("Blade Passing Frequency" or BPF), the frequency of "pure sound" at which individual blades pass through. a given fixed point in space. As a result, gas turbine engines are complex sources of noise, and the compressor rotor is a major noise source. In APUs and other wheel-type rotor applications, the noise is dominated by pure sound associated primarily with GMP, which exceeds the broadband noise portion of the acoustic spectrum. The noise intensity is also a function of aero-acoustic-mechanical interactions between the rotor blades and the driving fluid, multiple sounds occurring at a harmonic of the motor shaft frequency or engine order . Acoustic energy also moves and redistributes away from the GMP as the shock wave fronts propagate away from the rotor and into the far field, resulting in a multi-tone noise spectrum with a characteristic from his to "circular saw". A turbine apparatus comprises a rotor with a hub section and a plurality of blades. The hub section is defined about an axis of rotation, and 4 is divided into first and second circumferential sectors. The blades are attached to the hub section, extending radially outward. Blades in the first sector have a first angular spacing, and blades in the second sector have a second angular spacing. The first angular spacing is different from the second angular spacing, so that the blades are asymmetrical about the axis of rotation of the rotor.

Figure 1 is a cross-sectional view of a turbomachine with an asymmetric spacing of rotor blades. Figure 2A is a perspective view of a rotor with asymmetric blade spacing in an even number embodiment of blades. Figure 2B is a perspective view of a rotor with asymmetric blade spacing, in an odd number embodiment of blades. Figure 3 is a frequency spectrum graph for a turbomachine rotor. Figure 4A is a velocity field contour graph for a rotor with symmetrical blade spacing. Figure 4B is a velocity field contour graph for a rotor with asymmetric blade spacing. Figure 1 is a cross-sectional view of a turbomachine with an asymmetrical rotor blade spacing. In this particular embodiment, a turbomachine 10 comprises an auxiliary power unit with an inlet chamber 12, a trapezoidal tooth coupling or shaft 14, a work wheel 16 and a load height wheel 18. The wheel The work wheel 16 and the load height wheel 18 rotate together about a turbine axis A, 5 coupled by the trapezoidal gear coupling 14. The work wheel 16 and the load height wheel 18 have an asymmetrical spacing. blades for reducing pulsations and pressure vibrations, increasing the service life and reducing the ambient noise profile of the turbomachine 10. The inlet chamber 12 is divided into a load inlet 20 for the working wheel 16 (on the left in Fig. 1) and a load height input 22 for the load height wheel 18 (on the right, facing the work wheel 16). Air (or other compressible fluid) is drawn through the inlet 20 of the working wheel 16, which compresses the fluid for a bleed air system or a pneumatic reservoir, or for another fluid supply. compressed. The inlet 22 supplies air to the load height wheel 18, which compresses the air for the combustion chamber section of the turbomachine 10. A turbine rotor extracts rotational energy from the gas expanding combustion, driving the load height wheel 18 and the working wheel 16 through the shaft 14. According to the embodiment, the shaft 14 is also coupled with a gearbox, a generator or another accessory device. The wheels 16 and 18 rotate at high speed, generating shock waves at the frequency due to blade wakes. In order to reduce vibration and lower the characteristic circular saw noise output profile, the rotors 16 and 18 have an asymmetric blade spacing. This reduces the acoustic energy at the GMP, moving energy to other motor orders and frequencies. In two-rotor embodiments, the work wheel 16 and the load height wheel 18 have the same number of blades, and the work wheel 16 is synchronized or rotated on the trapezoidal gear coupling 14 with respect to the load height wheel 18. As a result, the work wheel blades rotate out of phase with the load height blades, further reducing the noise and vibration effects as described below. Figure 2A is a perspective view of a rotor for a turbomachine, with asymmetric blade spacing. In this particular embodiment, the rotor 30 includes a work wheel or load height wheel for an auxiliary power unit. In a variant, the rotor 30 comprises a wheel, a compressor rotor or a turbine rotor for an APU, a double flow reactor, a turboprop or turbine engine, or another blade turbine engine component such as a rotor for a pump. , a blower or a turbine.

The rotor 30 includes a frustoconical hub section 32 with main (long) blades 34 and alternately steered (short) vanes 36. Alternatively, the rotor 30 is called the rotor drum or disc, and the hub section 32 may have a generally conical, cylindrical or barrel shape.

In wheel and compressor embodiments, the rotor 30 is typically formed of a light and strong metal, such as titanium, or a similar alloy. In turbine embodiments, the rotor 30 is made of high temperature materials such as nickel, cobalt, and iron based alloys or superalloys. Alternatively, the rotor 30 is formed from a range of other materials, including aluminum, brass, and other metals, or from graphite, polymers, and composite materials. According to the embodiment, the blades 34 and the guide vanes 36 (if present) can be integrally formed with the hub section 32, for example by molding or machining a single cast or forged piece. Alternatively, the blades 34 and the guide vanes 36 are formed separately, and attached to the hub section 32 by welding, or using a mechanical attachment to the root.

As shown in FIG. 2A, the rotor 30 comprises an even number of main blades 34 and guide vanes 36, for example sixteen. The main blades 34 are circumferentially spaced around the rotor 30, extending radially from the hub section 32 and 8 axially along the axis of rotation A. The guide vanes 36 are between the main blades 34. The main blades 34 are divided into a first sector of 180 ° 38A comprising seven blades 34 with a first spacing between blades S1 25.7 °, and second sector of 180 ° 38B comprising nine blades 34 with a second spacing between blades S2 20, 0 °. The first blade spacing S1 is thus different from the second blade pitch S2, in order to reduce vibrations and to lower the circular saw noise profile by displacing the acoustic energy out of the frequency due to blade wakes. The blade spacing is an angular measurement, extending circumferentially about the axis A between the adjacent main blade centers or media 34, through the guide blades 36. In general, the angular separation remains constant, then that the actual separation distance between the blades 34 varies in the radial direction (i.e., moving away from the hub section 32, perpendicular to the axis of rotation). Alternatively, the angular spacings S1 and S2 differ essentially in the transonic and supersonic regions of the blades 34, where the shock waves are formed. In these embodiments, the angular spacings S1 and S2 are defined at a particular span height, for example a span height of 90 or at the blade tip (span height 100 δ). In these embodiments, the rotor 30 and the hub 32 are divided into two equal circumferential sectors 38A and 38B, with different blade numbers N1 and N2. Each sector occupies half of the rotor circle (i.e., 180 ° circumferential angle), corresponding to blade spacings S1 = 180 ° / N1 and S2 = 180 ° / N2, respectively.

For even numbers of N blades, the first sector 38A comprises N1 = N / 2 - 1 or less blades 34 (and guide blades 36), and the second sector 38B includes N2 = N / 2 + 1 or more blades 34 (and steering blades 36). The angular spacing is defined for each of the pairs of adjacent blades in the corresponding groups of blades. The blades on the periphery of each sector will have a larger spacing on one side (for example, towards the sector 38A), and a smaller blade spacing on the other side (for example, towards the sector 38B).

In these even-numbered embodiments of blades, the numbers of blades N1 and N2 in sectors 38A and 38B differ by an even number. In the nominal or first order embodiment, therefore, N1 = N / 2 - 1 and N2 = N / 2 + 1, with a blade difference of N2 - N1 = 2. In this embodiment, the spacing of blades is: 0 S1 (even) - 180 [1] N / 2-1 0 S2 (even) - 180 [2] N / 2 + 1

In other (higher order displacement) embodiments, the number of blades may differ by four, six or more, and the blade spacing therefore varies. It is also possible to divide even numbers of blades and vanes and vanes into two equal sets of N1 = N2 = N / 2 vanes each, and to define unequal circumferential sectors of different angular sizes. In these embodiments, one set of blades occupies more than 180 ° of the rotor circumference and the other set of blades occupies less than 180 °, and the blade spacing therefore varies. The guide vanes 36 are also divided into groups 38A and 38B. Typically, the spacing between the blades 34 and the adjacent guide blades 36 is half of the blade spacing, so the spacing between the guide vanes in each sector is substantially equal to the inter-blade spacings S1 and S2, respectively. Alternatively, the guide vanes 36 are displaced relative to the vanes 34, and the spacing between vanes and the spacing between the vanes varies. Non-uniform asymmetric blade spacing is contrary to general teachings and common sense in the art that blades and guide vanes must be arranged symmetrically to reduce imbalance. In general, the balance is maintained by machining or milling the hub section 32 to compensate for mass redistribution due to different blade spacings S1 and S2. Material may be added or removed, for example over an equilibrium range or an equilibrium circumference, or along the wheel bore or the rear face of the frustoconical portion of the hub section 32. Alternatively, the balance is maintained by drilling or forming apertures in the rotor 30, by adding or by adjusting equilibrium weights, or by false round or intentional shifting or by modifying the pitch diameter of trapezoidal teeth .

While the angular spacing between the blades varies, however, the blades 34 may be characterized by a substantially identical geometric profile. In particular, the blades 34 of Figure 2A all have substantially the same matrix of mass and stiffness, providing substantially the same eigenvalues, natural vibration frequencies and corresponding mode shapes. This is different from blade tuning techniques, in which individual blades have different geometric profiles, different matrices of mass and stiffness, and different eigenvalues and natural frequencies. FIG. 2B is a perspective view of a rotor variant 30 for a turbomachine, with asymmetrical blade spacing. The rotor 30 comprises a hub section 32, main blades 34 and guide blades 36. For odd numbers of blades N, as shown in FIG. 2B, the first sector 38A comprises (N-1) / 2 or fewer blades 34 (and guide blades 36), and the second sector 38B comprises (N + 1) / 2 or more blades 34 (and guide blades 36). In this case, the difference in the number of blades is odd. For the nominal or first-order case N1 = (N-1) / 2 and N2 = (N1 + 1) / 2 (ie, N2 - N1 = 1), the corresponding blade spacing is: 3600 S1 (odd ) =, [3] N-1 36001 S2 (odd) =. [4] In a particular embodiment, the rotor 30 comprises fifteen main blades 34 and fifteen guide blades 36, as shown in Figure 2B. The first sector 38A comprises seven blades 34 with the angular spacing S1 25.7 °, and the second sector 38B comprises eight blades 34 with the angular spacing S2 25.0 °. The first blade spacing S1 is thus greater than the second blade pitch S2, and the blades are asymmetrical around the rotor axis.

Asymmetric blade spacing reduces the level of rotor and blade vibration due to aeromechanical coupling, and lowers the sound intensity and environmental impact of the generated acoustic field. In particular, the asymmetric blade spacing reduces the wheel frequency noise at the GMP by reducing the Lighthill turbulence stress tensor, compared to a rotor with a symmetrical distribution of the same number of blades. Asymmetric blade spacing also reduces the amplitude of aerodynamic pressure oscillations experienced by diffuser vanes, maintaining blade integrity by reducing resonance effects at natural vibration frequencies. In particular, asymmetric blade spacing reduces the amplitude of pressure excitations through effects and 13 harmonics, reducing the likelihood of damage or failure due to high cycle fatigue (HCF). ), which can occur when cyclic stresses exceed the endurance limits of the blade material. The vibration effects are further reduced by phase shifted synchronization of multiple rotors, as described further below. Fig. 3 is a frequency spectrum graph for a turbine rotor. The frequency is given along the horizontal axis, with the amplitude along the vertical, both in arbitrary units. Frequency peaks are indicated by motor order (FO).

The sources of rotor vibration can be separated into two main categories or types: mechanical and aerodynamic. Mechanical causes are related to the dynamic characteristics of the rotor, and are influenced by a number of factors, including unbalance, misalignment of bearing support, false shaft, and damaged or rubbed parts. The aerodynamic causes are related to pressure pulsations at multiples distinct from the rotor rotation frequency and the number of blades (ie, multiples of the GMP). The aerodynamic sources arise from perturbations generated by the blades 34, which are transmitted to the rotor 30. In the rotor frequency spectrum, the first motor order (1EO) represents the rotor imbalance effect. The second, third, fourth, sixth, eighth and other harmonics are due to mechanical factors, as described above. A rotor or a wheel with sixteen main blades, for example, will have dynamic effects characterized by the second (2E0),

fourth (4E0) and eighth (8E0) engine orders, because of the common factors of two, four and eight. A rotor or a wheel with fifteen main blades will present dynamic effects characterized by the third (3E0) and fifth (5E0) engine orders, in

because of the common factors of three and five.

The acoustic field is described by an inhomogeneous wave equation, which can be derived from the Navier-Stokes equations. Using Einstein's notation: 2 at ~ - co2V2p = ~ [5] where p is the density of air or fluid, co is the velocity of ambient noise and Ti7 is the Lighthill turbulence stress tensor. The generation of

its (noise) is attributed to the right side of this equation.

The Lighthill turbulence stress tensor is further defined by: 2 Tu = pu; uj - at + (p - co p) 6a, with velocity components vi and vi, a viscous stress contribution oij and a pressure p . The term pviv is the Reynolds stress tensor, representing the irregular flow convention. The viscous stress tensor oij

represents the noise generated by shear, and the [6] pressure p is incorporated into the nonlinear acoustic or acoustic term (p-co2p), the Kronecker delta 5i being defined to be one for i = j and zero for i j. The blade turbine components generate pressure pulsations with a distinct strong component at the frequency due to the blade wakes, for example a GMP corresponding to 15E0 or 16E0, depending on the number of blades. For a rotor having dynamic characteristics expressed in harmonics of the motor order, the amplitude of BPF will generally influence the rotor stability. In a rotor with asymmetric blade spacing, the periodic and cyclic pressure field characteristics are reduced by the introduction of two blade sectors having different numbers of blades N1 and N2, with different angular spacings S1 and S2. The pressure function is expressed in terms of a Fourier series in the BPF frequency: P (t) = Po + a ~ sin (lEOxt + al) + ... + aNsin [(N) EOxt + aN) +. .. [7] The coefficients an represent the relative amplitudes of the sinusoidal function at each motor order, and at a relative phase an. The index n goes from n = 1 for the first motor order 1EO, to harmonics of order n = N and more. In an odd-numbered asymmetric configuration with N1 = (N-1) / 2 blades at spacing S1 and N2 = (N + 1) / 2 blades at spacing S2, there are significant contributions to motor 2Nl = N - 1 and 2N2 = N + 1: P (t) = Po + b ~ sin (lEO xt + / 31) + ... + bN sin [(N -1) EO xt + NN-1] [8] + bN sin [(N) EO xt + NN] + bN + 1 sin [(N + 1) EO xt + NN + 1] + ... In particular, the asymmetric configuration has the following amplitude characteristics aerodynamic pressure: 1 aN 1 (asymmetrical) <1 + £ bN ~ (symmetrical), [9A] (asymmetrical) <1 bN (symmetrical), [9B] 1 + s 1 aN + 1 (asymmetric) <1 + £ bN + 1 (symmetrical). [9C] The difference or magnitude reduction coefficient (s) is about thirty percent (30%) or more (i.e., s = 0.30). Conversely, therefore, unmodified (symmetrical) amplitudes bN-1, bN and bN + 1 are at least 30 δ greater than the (asymmetric) amplitudes aN-1, aN and aN + 1.

For the case of a fifteen-blade rotor (fifteen steer blades) divided into two groups of seven and eight blades, respectively, seven-blade sector generates a 25 BPF at engine order 14E0 (at the second sector of eight blades generates a BPF pressure signal at motor order 16E0 (i.e., 2 x 8) This gives: and signal to know, 2 or sectors the first pressure of x 7), and the a14 <1 xb14, [ 10A] 1 + s a15 <1 x b15, [10B] 1 + s 1 a16 <1 + £ Xb16. [1OC] The reduction factor is at least 30% in these motor orders, as described above. For even numbers of N blades, the rotor is

10 divided into a first sector with N1 = N / 2 - 1 blades at the spacing S1, and a second sector with N2 = N / 2 + 1 blades at the spacing S2. The asymmetric contributions appear at the orders of motor (N-2) EO and (N + 2) EO: P (t) = Po + bl sin (lEO xt + / 31) + ... + bN 2 sin [(N - 1) EO xt + NN2] + ... [11] + bN sin [(N) EO xt + NN] + bN + 2 sin [(N + 1) EO xt + PN + 2] + ... The reduction of amplitude is again at least 30 giving: 1 20 aN2 (asymmetric) <1 + bN 2 (symmetrical), [12A] (asymmetric) <1 bN (symmetrical), [12B]

1 + s 1 and 15 and aN + z (asymmetric) <1 + £ bN + 2 (symmetric). [12C] For a sixteen blade rotor divided into sectors of seven and nine blades, respectively, the first sector generates an excitation BPF at the order of 14E0 motor (2 x 7 blades), and the second sector generates a GMP excitation at engine order 18E0 (2 x 9 blades). So . a14 <1 xbt4, [13A] 1 + £ a16 <1 xb16, [13B] 1 + £ and 1 als <+ sxblg. 1 [13C] In a turbomachine with two or more rotors, each rotor contributes to the total pressure amplitude. To jointly rotate the charge and load height wheels having the same number of blades N, for example, the net pressure signal is: P (t) = Po + and sin (lEO xt + / 31) + .. . + cN sin [(N) EO xt + NN] + ... [14] + dl sin (lEO xt + yl) + ... + dN sin [(N) EO xt + yN] + ...

The primary motor sequence (first motor order) 1EO is defined by the frequency due to blade wakes, which depends on the number of blades N and the angular frequency w: 1EOBPF = Nw. [15] For asymmetric blade spacing at odd numbers of N blades, harmonic "sideband" contributions occur at motor (N ± 1) EO orders. For even numbers of 19 N blades, contributions appear at motor orders (N ± 2) EO. Frequency-compatible rotors have the same number of blades N, divided into similar sets of blades N1 and N2 with similar blade spacings S1 and S2, giving the same excitation frequencies BPF. In order to move the two contributions out of phase, a rotor is synchronized or rotated at an angle of 360 ° / 2N around the motor axis. This gives: yn = / 3n + 7c. [16] In general, two signals of the same frequency with the same counterbalancing (or counterpropagating) amplitude may be canceled or negated by placing them out of phase (or in destructive interference). The amplitude or quantity of the phase shift is a function of wavelength and the distance between two sources. Mechanically, the phase shift is accomplished through the shaft or linkage linking the two wheels to the rotor assembly. For trapezoidal gear couplings, for example, trapezoidal "tooth" structures (notches or notches) are typically aligned at each end, and the annular offset is accomplished by rotating one of the wheel rotors relative to the teeth. In a particular embodiment, the trapezoidal gear coupling has 2N notches or trapezoidal teeth (i.e., twice the number of blades). In this embodiment, one of the wheels is rotated by a tooth or notch, producing an angular shift of 360 ° / 2N to generate the required phase shift. Alternatively, the wheel is rotated by an odd number of teeth. In principle, rotors with equal charges will have substantially similar amplitude coefficients in and dn, and the net pressure field may be closer to zero at some points. In practice, however, the rotors are not always of the same size, and do not undergo the same loading, so the amplitudes are different. In addition, the relative phase shift varies by location, generating a stationary wave pattern in (at least) the near and intermediate fields. These effects are addressed by adjusting tip free space, crankcase eccentricity, false rotor circumference, and blade stiffness to produce rotors with substantially equivalent impedance. This further minimizes the net pressure signal, including the far field region where environmental effects are a problem. Fig. 4A is a speed field graph for the rotor 30 with symmetrical spacing of the blades 34 and guide vanes 36 around the hub 32, and Fig. 4B is a graph for asymmetric spacing. The two graphs are taken at a blade span of 95 Mach outlines being represented in arbitrary units. FZ flow separation zones correspond to low Mach numbers of about 0.30 M or 0.15 M, and are associated with an acoustic source intensity. Comparing Figure 4A with

Figure 4B, the flow separation for the asymmetric blade spacing (Figure 4B) is smaller than that for the symmetrical case (Figure 4A), and the size of the flow separation zones FZ is reduced. Thus, for the same number of blades, the asymmetric blade spacing (FIG. 4B) causes a relatively lower sound intensity than that of the symmetrical blade spacing (FIG. 4A). Although the present invention has been described with reference to exemplary embodiments, those skilled in the art will understand that various changes may be made and equivalents may replace elements thereof without departing from the mind and of the scope of the invention. In addition, modifications may be made to adapt a particular situation or material to the teachings of the invention, without departing from the essential scope thereof. Therefore, the invention is not limited to the particular embodiments described herein.

Claims (20)

REVENDICATIONS1. A turbine apparatus comprising: a rotor (30) having a hub (32) defined about an axis of rotation (A), the hub (32) having first and second circumferential sectors (38A, 38B); and a plurality of blades attached to and radially extending from the hub (32) therefrom, the plurality of blades including a first group of blades having a first angular spacing (Sl) in the first sector (38A) and a second group of blades having a second angular spacing (S2) in the second sector (38B); wherein the first angular spacing (Sl) is different from the second angular spacing (S2), and the rotor blades are asymmetrical about the axis of rotation (A). The turbine apparatus according to claim 1, wherein the plurality of vanes comprise an odd number N of vanes, the first group of vanes comprising (N-1) / 2 of the N vanes and the second group of vanes comprising (N + 1) / 2 of the N blades. The turbine apparatus according to claim 2, wherein the first angular spacing (Sl) is approximately 360 ° / (N-1) and the second angular spacing (S2) is approximately 360 ° / (N + 1). The turbine apparatus according to claim 1, wherein the plurality of blades comprises an even number N of blades, the first group of blades comprising 23 N / 2-1 of the N blades and the second group of blades comprising N / 2 + 1 of the N blades. A turbine apparatus according to claim 4, wherein the first angular spacing (Sl) is approximately 180 ° / (N / 2-1) and the second angular spacing (S2) is approximately 180 ° / (N / 2 + 1) ). The turbine apparatus of claim 1, further comprising a plurality of guide vanes (36) attached to the hub section (32) and disposed between the plurality of vanes. A turbine apparatus according to claim 6, wherein the plurality of guide vanes (36) comprises a first group of guide vanes having the first angular spacing (Sl) in the first circumferential section and a second set of guide vanes comprising the second angular spacing (S2) in the second circumferential section. The turbine apparatus according to claim 1, wherein the rotor (30) comprises a wheel and further comprising a compressor rotor rotatably coupled to the wheel, the compressor rotor comprising: a compressor hub defined around the wheel; rotational axis, the compressor hub having first and second circumferential sectors; a second plurality of blades attached to and extending radially from the second hub, the second plurality of blades including a first group of blades having the first angular spacing (Sl) in the first sector of the second hub 24 and a second group of blades having the second angular spacing in the second sector of the second hub. An auxiliary power unit comprising the turbine apparatus according to claim 8, and further comprising a trapezoidal gear coupling (14) for rotatably coupling the second rotor with the first rotor. An auxiliary power unit comprising the turbine apparatus according to claim 8, wherein the first rotor is synchronized with the second rotor such that the first plurality of blades is out of phase with the second plurality of blades. A wheel, comprising: a hub section defined about an axis of rotation, the hub section including first and second circumferential sectors; a first set of blades extending radially from the first sector of the hub section, the first set of blades having a first angular spacing; and a second set of blades extending radially from the second sector of the hub section, the second set of blades having a second angular spacing; wherein the first angular spacing is different from the second angular spacing, so that the first and second sets of blades are asymmetrical about the axis of rotation. 25 Wheel according to claim 11, wherein the first set of blades comprises an odd number of blades N1 and the first angular spacing is about 180 ° / Nl. Wheel according to claim 12, wherein the second set of blades comprises an even number of blades N1 + 1 and the second angular spacing is about 180 ° / (Nl + 1). Wheel according to claim 12, wherein the second set of blades comprises an odd number of blades N1 + 2 and the second angular spacing is about 180 ° / (N1 + 2). 15. Auxiliary power unit comprising the wheel according to claim 11. 16. The auxiliary power unit of claim 15, further comprising a compressor rotor rotatably coupled to the wheel, the compressor rotor comprising: a compressor hub defined about the axis of rotation, the compressor hub; comprising first and second circumferential sectors; a third set of blades extending radially from the first sector of the compressor hub, the third set of blades having the first angular spacing; a fourth set of blades extending radially from the second sector of the compressor hub, the fourth set of blades having the second angular spacing. 30 A turbo machine (10), comprising: a work wheel (20) circumferentially divided into first and second sectors (38A, 38B), the first sector (38A) of the work wheel (20) having a set of blades with a first angular spacing (Sl) and the second sector (38B) of the working wheel (20) having a set of blades with a second angular spacing (S2); a load height wheel (18) circumferentially divided into first and second sectors (38A, 38B), the first sector (38A) of the load height wheel (18) having a set of blades with the first angular spacing (S1) and the second sector (38B) of the load height wheel (18) having a set of blades with the second angular spacing (S2); and a rotatable coupling (14) between the work wheel (20) and the load height wheel (18), the coupling defining an axis of rotation (A); wherein the first angular spacing (Sl) and the second angular spacing (S2) are different, so that the working wheel (20) and the load height wheel (18) are asymmetric about the axis of rotation ( AT). The turbomachine (10) according to claim 17, wherein the working wheel (20) is synchronized about the axis of rotation (A) with respect to the load height wheel (18) so that the wheel working shaft (20) is out of phase with respect to the load height wheel (18). The turbomachine (10) according to claim 17, wherein the work wheel (20) and the load height wheel (18) each comprise an equal odd number of blades N, and wherein the first angular spacing (S1) ) is about 360 ° / (N-1) and the second angular spacing (S2) is about 360 ° / (N + 1). The turbomachine (10) according to claim 17, wherein the work wheel (20) and the load height wheel (18) each comprise an equal even number of blades N, and wherein the first angular spacing (Sl) is about 180 ° / (N / 2-1) and the second angular spacing (S2) is about 180 ° / (N / 2 + 1).