RADIAL IMPELLER FOR A DRUM FAN AND FAN UNIT HAVING A RADIAL IMPELLER OF THIS TYPE

Abstract

An impeller has an inlet and outlet. A bottom disc has an external diameter. A top disc concentric to the bottom disc at an axial distance therefrom has a suction opening for the inlet. A plurality of forward-curved blades are arranged between the bottom disc and the top disc. A flow channel has an inner inlet side and an outer outlet side formed between adjacent blades. The channel is curved convexly viewed in the running direction of the impeller. The external diameter of the bottom disc is at least 20% greater than the internal diameter of the suction opening. The top disc forms a guiding surface and an angle is formed between a tangent on the guiding surface at the inlet to the suction opening and a tangent on the guiding surface at the outlet from the flow channel on the air outlet side is at least 30°.

Description

FIELD

The present disclosure relates to a radial impeller having an axial air inlet and air outlet over the circumference of the impeller, preferably for use in a spiral-shaped housing, comprising a bottom disc which has an external diameter, and comprising a top disc which is arranged concentrically to the bottom disc at an axial distance therefrom and has a circular suction opening with an internal diameter for the axial air inlet, and also comprising a plurality of forward-curved, profiled blades arranged between the bottom disc and the top disc, a flow channel having a radially inner air inlet side and a radially outer air outlet side being formed between two adjacent blades in each case, which channel is curved convexly, viewed in the running direction of the radial impeller. This channel curvature means that the suction-side surfaces of the blades are curved in a convex manner, at least in regions, and their pressure-side surfaces are curved in a concave manner, at least in regions.

The disclosure also relates to a fan unit, in particular to a drum fan, having a radial impeller of this type. The term “fan” subsumes fans having a pressure ratio between suction side and pressure side within a range of 1.0 to 1.1, as well as fans having a pressure ratio between suction side and pressure side within a range of 1.1 to 3.0.

BACKGROUND

This section provides background information related to the present disclosure which is not necessarily prior art.

Nowadays, radial impellers are preferably used where a high build-up of pressure is to be achieved with a relatively low volume flow. Since in the case of radial impellers, the entire conveyed air flow leaves the impeller at the external diameter, it is possible to generate a greater kinetic energy of the air molecules and thereby also a higher pressure compared to an axial fan, the circumferential speed of which is restricted at the wheel hub. The use of radial impellers is particularly effective when an air flow has to be diverted by 90° from the axial direction into the radial direction, or when components, filters or the like obstruct a free air flow. The most common configuration is that of a complete radial fan with a housing, although there are also different motor-impeller combinations for uses in which the air conveyance for the build-up of pressure can be integrated into one device.

In the case of radial impellers, a distinction is made between impellers with blades which are curved forwards in the running direction, those with backward-curved blades and those with radially ending blades. Blades which are curved forwards, in the direction of rotation, cause the flow channel which runs from a radially inner air inlet side to a radially outer air outlet side, to be curved in a convex manner, viewed in the running direction of the radial impeller. The suction-side surfaces of the blades are thus curved in a convex manner, at least in regions, and their pressure-side surfaces are curved in a concave manner, at least in regions. Radial impellers having forward-curved blades allow a high angular momentum to be delivered to the air flow and thus achieve a high energy conversion. However, a disadvantage here is a high dynamic pressure of the outgoing air. This dynamic pressure has to be converted into static pressure in a subsequent guiding apparatus, for example in a spiral housing. Radial impellers with forward-curved blades deliver more angular momentum to the flow than radial impellers with backward-curved blades. Thus, the necessary speed for reaching the same operating point in radial impellers with forward-curved blades is substantially lower compared to radial impellers with backward-curved blades of the same size. The efficiency of radial impellers having backward-curved blades is significantly higher compared to radial impellers having forward-curved blades.

A particular configuration of the radial fan is the drum fan. Designated as drum fans are radial fans, the impellers of which are identical to a drum, i.e. the width of the impeller is relatively large compared to the diameter thereof. In particular, it can be within a range of from 40 to 80 percent, based on the external diameter of the bottom disc. Rotors of this type, provided with forward-curved blades and also known for about 80 years under the name of Scirocco rotors, are used where small radial dimensions are required. The ratio of the internal diameter of the top disc to the external diameter of the bottom disc of the original Scirocco rotor was 0.875.

Present-day forward-curved radial impellers which are used in spiral housings as cylindrical rotors are distinguished by a high power density, i.e. by a high conveyed volume with a small installation space and by good acoustic characteristics, in particular by a low noise level during operation. However, the aerodynamic efficiency is relatively low compared to backward-curved impellers due to burbling and vortex formation. Cylindrical rotor fans are used in ventilation and air-conditioning technology in installations requiring a pressure increase of preferably up to 4000 Pa and with volume flows of up to 8 m³/s, based on an impeller diameter of one metre and a single-flow configuration.

In a radial impeller known, for example, from DE 10 2006 031 167 A1 of the type mentioned at the outset, having an axial air inlet and air outlet over the circumference of the impeller, a strong profiling of the blades can prevent or minimise burbling in the flow channel between the blades. In this respect, “profiling” means that the thickness of the blades varies over the extension direction thereof, the blades more particularly then being considered as strongly profiled when a so-called profile thickness ratio therein, i.e. a ratio of profile thickness to overall profile length, is more than or equal to 0.15, in particular more than or equal to 0.2 and more preferably more than or equal to 0.25. In this respect, the profile thickness ratio is preferably at most 0.5, in particular 0.4 and more preferably 0.35. The top disc not shown in the figures of the mentioned document is called a frame in said document and is part of the spiral housing. Burbling at the top disc is disadvantageously reinforced by the known blade profiling. Thus, this measure can only very slightly increase the efficiency.

SUMMARY

This section provides a general summary of the disclosure, and is not a comprehensive disclosure of its full scope or all of its features.

The object of the present disclosure is to provide a radial impeller of the type mentioned at the outset and a fan unit having a radial impeller of this type, in which the advantages of a cylindrical rotor are combined with the advantages of an impeller of this type having backward-curved blades, thus with which it is possible to achieve in particular an increase in efficiency while retaining a high power density and a low development of noise.

This is achieved according to the disclosure in that the external diameter of the bottom disc of the radial impeller, based on the internal diameter of the suction opening, is at least 20% greater than the internal diameter of the suction opening, in that the top disc forms a top guiding surface in the flow channel, an angle which opens out towards the inlet opening and is formed between a tangent on the top guiding surface at the inlet to the suction opening of the top disc and a tangent on the top guiding surface at the outlet from the flow channel on the radially outer air outlet side thereof, being at least 30°.

A fan unit according to the disclosure is distinguished in that the radial impeller according to the disclosure is arranged in a housing, particularly in a spiral housing.

Due to the disclosure, it is possible to retain the advantage of radial impellers known hitherto as cylindrical rotors, in that these have a high power density and a low noise with slight rotational noise increases compared to radial impellers having backward-curved blades. The advantageously high power density can be attributed to a high angular momentum delivered to the flow by the forwards curvature of the blades. In this respect, a low development of noise is promoted by preferably high numbers of blades and preferably low speeds during operation. Both the forwards curvature and particularly also a large number of blades prevent or reduce at least burbling at the blades and at the top disc, but they increase the friction forces, which can lead to losses and to a reduction in efficiency. This can be effectively counteracted by the geometric configuration according to the disclosure of the radial impeller, and a preferred configuration can provide that the external diameter of the bottom disc, based on the internal diameter of the suction opening of the top disc, is greater by 50%, at most by 90%, than the internal diameter of the top disc.

The configuration according to the disclosure of the diversion of the axial onflow into a radial or diagonal direction can prevent in particular burbling at the top disc, which can be detected by a so-called CFD flow simulation. CFD (computational fluid dynamics) denotes a method which is established in flow mechanics and has the objective of solving fluid-mechanical problems iteratively by numerical methods and then visualising the results, preferably by colour representations. In this respect, the Navier-Stokes equations used in fluid mechanics and describing momentum conservation and mass conservation are modelled mathematically while presetting specific marginal conditions. This is an economical alternative to expensive experimental test series which are carried out, for example in a wind tunnel, and it makes it possible to analyse flow parameters which cannot be determined by measurements, such as turbulent kinetic energy, vortex viscosity, etc.

To prevent burbling at the top disc, a configuration of the top disc with a preferably relatively great axial width is also significant, according to which configuration the width of the top disc can preferably occupy at least 30% of the overall width of the impeller. A width configuration of this type is to be considered as being synergistically effective in combination with the disclosure, since in the case of conventional cylindrical rotors with unprofiled blades, an improvement cannot be achieved by this configuration.

With a predetermined shaft output, the efficiency is determined by the conveyed volume flow and by the total pressure increase caused by the fan, the product of which gives the conveying capacity, the term “total pressure” being understood as meaning the sum of static and dynamic pressure according to the so-called Bernoulli equation. Thus, the efficiency describes the ratio of conveying capacity to power of a shaft driving the fan and is calculated according to the formula:

η=(V*Δp_t)/P_W

where η denotes the non-dimensional efficiency, V denotes the volume flow in m³/s, Δp_t denotes the total pressure increase in Pa and P_W denotes the shaft output in W.

The combination of the “profiled blades” feature with the features of the top disc configured in terms of flow according to the disclosure can significantly increase the efficiency. However, this is only possible when the external diameter of the bottom disc is at least 20% greater than the internal diameter, determining the size of the suction opening, of the top disc, in contrast to the previous known embodiments.

On the other hand however, with unprofiled blades, as are also known from the prior art, it is not possible, even with the diameter ratio provided according to the disclosure and the minimum angle between the tangents of 30°, to divert the flow in the blade channel without any burbling. As is known, burbling involves losses and results in poor impeller efficiency. Thus, as is known, it is even assumed that an imperfect onflow at the blade inlet due to excessively steep blades is still to be considered as more favourable than an unstable burbling which occurs in a blade-congruent flow. Due to a blade profiling, the flow can be deflected in the blade channel without any burbling in the configuration, provided according to the disclosure, of top disc and bottom disc, in spite of a blade-congruent onflow, which implies low impact losses at the blade inlet.

It is possible to further increase the efficiency of an entire fan with the radial impeller according to the disclosure by selecting a width ratio of a housing width at the impeller outlet to the impeller outlet width itself of at least 1.0 to at most 1.4.

Thus, with the disclosure it is possible to achieve efficiencies within a range of between 0.65 and 0.80, preferably even up to 0.90.

When a backward-curved radial fan according to the prior art is operated at the optimal acoustic operating point, i.e. at maximum efficiency, it is possible to estimate the total sound power level with an accuracy of ±4 dB according to the formula:

L_W=37+10 log(V)+20*log(Δp_t).

In this formula, L_W denotes the total sound power level in dB, V denotes the volume flow in m³/s and Δp_t denotes the total pressure increase in Pa. However, this formula cannot be applied to the radial impeller according to the disclosure. Compared to the measured values or to the values which are calculated according to the above L_W formula, for radial impellers according to the prior art, a radial impeller according to the disclosure, for example with an external diameter of 170 mm achieves an improvement of more than 4 dB in the acoustically optimised operating range.

The coefficient of performance L which is to be considered as an indication of the power density is understood, according to the formula:

L=φ_r*ψ

as the product of capacity coefficient φ_r and coefficient of pressure ψ. In this respect, all the quantities are non-dimensional, the capacity coefficient φ_r being calculated according to the formula:

φ_r=(V*60)/(b*π²*D²*n)

and describing the ratio of the actual conveyed quantity to the theoretically possible conveyed quantity. The capacity coefficient results from the product of the outlet surface of the wheel and of the circumferential speed. In the formula, (Pr is the capacity coefficient, the index r representing “radial”, V is again the volume flow in m³/s, D is the impeller external diameter in m which is determined by the outlet diameter D_a,S of the blades, b is the outlet width of the impeller in m and n is the speed in 1/min.

The coefficient of pressure is the ratio of the pressure level, generated by the wheel, to the dynamic pressure of the circumferential speed and is calculated according to the formula:

Ψ=(Δp_t*2*60²)/(p*(D*π*n)²)

where Ψ is the non-dimensional coefficient of pressure, p is the density in kg/m³, Δp_t is the total pressure increase in Pa, D is again the impeller external diameter in m determined by the outlet diameter D_a,S of the blades, and n is the speed in 1/min.

With the disclosure, it is possible to achieve capacity coefficients within a range of 0.6 to 1.0, preferably within a range of 0.6 to 0.8 and coefficients of pressure within a range of 2.2 to 3.2, preferably within a range of 2.8 to 3.0, it being possible for the coefficient of performance to be within a range of 0 to 1.5, preferably within a range of 0 to 1.0.

Unlike conventional cylindrical rotors in which, when the impeller is installed, there is usually an axial distance in the region of a few millimetres between nozzle and top disc; in the configuration according to the disclosure of the top disc, it is also optionally advantageously possible to provide a nozzle-shaped configuration of the inlet and an axial dipping of the nozzle into the top disc. Consequently, the split flow can be aimed in the same direction as the main volume flow entering through the suction opening. The split flow then advantageously contributes to the stabilisation of the deflection into the radial direction, which, as is known, only happens in the case of radial wheels with backward-curved blades.

Further areas of applicability will become apparent from the description provided herein. The description and specific examples in this summary are intended for purposes of illustration only and are not intended to limit the scope of the present disclosure.

DRAWINGS

The drawings described herein are for illustrative purposes only of selected embodiments and not all possible implementations, and are not intended to limit the scope of the present disclosure.

Advantageous configurations of the disclosure are contained in the sub-claims and will be described in more detail with reference to the embodiments shown in the accompanying drawings, in which:

FIG. 1 is an axial part-sectional view of a preferred configuration of a radial impeller according to the disclosure,

FIG. 2 is a view of a cross section along line II-II in FIG. 1, of the configuration of the radial impeller according to the disclosure shown in FIG. 1.

FIG. 3 is similar to FIG. 1, but is a full sectional view of the configuration of a radial impeller according to the disclosure installed in a fan unit according to the disclosure,

FIG. 4 is an axial half-sectional view of a second configuration of a radial impeller according to the disclosure in a fan unit according to the disclosure,

FIG. 5 is a view, like that of FIG. 4, and simplified compared to FIG. 3, of the first configuration of a radial impeller according to the disclosure in a fan unit according to the disclosure,

FIG. 6 is a view, like that of FIG. 1, of a second configuration of a radial impeller according to the disclosure.

Corresponding reference numerals indicate corresponding parts throughout the several views of the drawings.

DETAILED DESCRIPTION

Example embodiments will now be described more fully with reference to the accompanying drawings.

In the figures of the drawings, identical parts or functionally identical parts are denoted by the same reference numerals and signs. However, if specific, described features and/or features which can be inferred from the drawings of the radial impeller or fan unit according to the disclosure or the constituents thereof are mentioned only in connection with the embodiments, they are also, according to the disclosure and independently of this embodiment, significant as individual features or also combined with other features of the embodiment and can be claimed as belonging to the disclosure.

As can firstly be seen from FIGS. 1 and 2, a radial impeller 10 according to the disclosure has a top disc 1, a plurality of forward-curved, profiled blades 2 and a bottom disc 3. The top disc 1 forms a suction orifice and thus has a circular suction opening 4 with an internal diameter D_i,DS for an axial air inlet. The bottom disc 3 has an external diameter D_a,BS and is arranged concentrically to the top disc 1 at an axial distance therefrom.

The blades 2 are located between the top disc 1 and the bottom disc 3. Formed between two blades 2 in each case is a flow channel 5 which, viewed in the running direction LR of the radial impeller 10, is curved convexly and in which the flow moves in a direction S from a radially inner air inlet side 5a to a radially outer air outlet side 5b. The curvature of the flow channel 5, which is convex at least in part, viewed in the running direction LR of the radial impeller 10, means that, as shown in FIG. 2, a pressure side 2c of the blades 2 which, viewed in the running direction LR of the radial impeller 10, is respectively located under the blades 2, is curved in a concave manner, at least in part, and a suction side 2d of the blades 2 which, viewed in the running direction LR of the radial impeller 10, is respectively located on the blades 2, is curved in a convex manner, at least in part.

The top disc 1, the blades 2 and the bottom disc 3 can preferably be configured as a composite body consisting of two parts, in particular of two plastics injection-moulded parts which are integrally bonded.

Profiling of the blades 2 means that the profile thickness d_S of the blades 2 is not constant over the length thereof. In this respect, characteristic of the profiling of the blades 2 is a profile thickness ratio which is described by the ratio V_P of maximum profile thickness d_S to profile overall length L_S (see in this respect FIG. 2) and which should be at least 0.15, in particular at least 0.2, and more preferably at least 0.25, it being possible for the profile thickness ratio V_P to be at most 0.5, in particular 0.4 and more preferably 0.35. The position of the maximum profile thickness d_S can preferably be within a range of 5% to 75% of the profile overall length L_S, seen from the air inlet side 5a, and from there is reduced both towards the leading edge 2a of the blades 2 and towards the trailing edge 2b. The advantages described above come into effect as a result of this particularly streamlined profiling, but according to the disclosure without any burbling phenomena of the flow occurring on the top disc 1.

To form a shape which is favourable in terms of fluid-mechanics, it can be optionally provided, as shown in FIG. 2, that the leading edges 2a and/or the trailing edge 2b of the blades 2 are rounded in each case. Further features which optionally or preferably describe the shape of the blades are a crescent-shaped, but asymmetrical cross section of the blades 2, an outer curvature, which is convex at least in part, of the suction side 2d which is greater than the inner curvature, which is concave at least in part, of the pressure side 2c, and a drop shape in respect of the curved centre axis through the blades 2.

An optimum number of blades 2, which number is characteristically large for a drum fan, is at least 19 and at most 54 and is preferably within a range of 22 to 46. High numbers of blades can block the flow channel 5 and reduce the maximum possible volume flow V. In addition, the friction losses on the blade walls can increase so that the efficiency η decreases.

Furthermore, as can be seen in FIG. 3 and also in FIGS. 4 and 5, a radial impeller 10 according to the disclosure is preferably intended for use in a fan unit 20 according to the disclosure. The radial impeller 10 according to the disclosure can be arranged in this fan unit 20 according to the disclosure, preferably coaxially with an electric drive motor 6 and it is positioned in a housing 7, which can preferably be a spiral-shaped housing 7, as shown in FIG. 3.

In the illustrated configuration, the fan unit 20 according to the disclosure is a fan having a forward-curved radial wheel. It can preferably be a drum fan, a characteristic of which is also that the overall width b_ges of the radial impeller 10 is within a range of 25% to 70%, based on the external diameter D_a,BS of the bottom disc 3. As shown in FIG. 1, the overall width b_ges is the sum of a width b_DS of the top disc 1 and of an impeller outlet width b₂ at the radially outer air outlet side 5b of the flow channel 5.

According to the disclosure, it is provided that the external diameter D_a,BS of the bottom disc 3, based on the internal diameter D_i,DS of the top disc 1, is at least 20%, preferably at least 50%, greater than the internal diameter D_i,DS of the top disc 1. The top disc 1 forms a top guiding surface 8 for the flow channel 5, an angle α_DS which opens out towards the inlet opening 4 and is formed between a tangent T₁ on the top guiding surface 8 at the inlet to the suction opening 4 of the top disc 1 and a tangent T₂ on the top guiding surface 8 at the outlet from the flow channel 5 on its radially outer air outlet side 5b, being at least 30°. The maximum value of this angle can be 90°, preferably 75°. In the first configuration, the tangent T₁ runs parallel to the longitudinal axis X-X of the radial impeller 1. In this way, according to the disclosure the flow is diverted in such a favourable aerodynamic manner from the axial direction into a radial or diagonal direction that the efficiency η increases while the advantages of conventional cylindrical rotors are retained.

This is also the case when the tangent T₁ deviates from the parallel course to the longitudinal axis X-X of the radial impeller 1 by an angular value α_DS1 of up to ±30°, but preferably by only up to ±5°, as shown by the second configuration according to FIG. 6. An angle which opens out towards the inlet opening 4 and is formed between the tangent T₂ on the top guiding surface 8 at the outlet from the flow channel 5 on its radially outer air outlet side 5b and the longitudinal axis X-X of the radial impeller 10 is denoted by reference sign α_DS2. Thus, the following equation applies to the angle α_DS claimed according to the disclosure:

α_DS=α_DS2−α_DS1.

Analogously to the guiding surface 8 on the top disc 1, the bottom disc 3 can also form a bottom guiding surface 9 in the flow channel 5.

The top guiding surface 8 and/or the bottom guiding surface 9 of the flow channel 5 can particularly be constant in curvature, as shown in the drawing, except in the case of the bottom disc 3 of the configuration in FIG. 4, which advantageously counteracts the formation of flow turbulence.

Instead of the above-mentioned distance, measured axially in the direction of the longitudinal axis X-X, between bottom disc 3 and top disc 1, FIGS. 1, 2 and 5 show in the flow channel 5, a shortest distance in each case, respectively designated by reference sign A, which preferably varies between air inlet side 5a and air outlet side 5b of the flow channel 5, between bottom disc 3 and top disc 1. It can advantageously be provided that this distance A decreases in direction S from the radially inner air inlet side 5a to the radially outer air outlet side 5b, particularly while considering the blade spacing determined by the number of blades 2 such that the cross section of the respective flow channel 5 also tapers. This is shown in particular in FIG. 5 in which this preferred configuration is compared with a hypothetical channel configuration which is indicated by a dash-dot line and for which this distance A is constant. As indicated in the drawing by the word “constant”, in the hypothetical configuration, which, although possible in the context of the disclosure, is not preferred, the top guiding surface 8 and the bottom guiding surface 9 run at an equal distance from one another.

As far as the installation is concerned in a fan unit 20 according to the disclosure, of a radial impeller 10 according to the disclosure, shown by way of example in FIGS. 3 to 5, various technical measures relating to the nature of this installation can optionally further advantageously contribute to the achievement of an increase in the efficiency η while retaining a high power density L and a low total sound power level L_W.

Thus, it can be provided in particular that an inlet 21 of the housing 7 for the radial impeller 10 is nozzle-shaped, the inlet 21 of the housing 7 dipping in particular into the suction opening 4 in the top disc 1, as shown most clearly in FIG. 3, but also in FIGS. 4 and 5.

This configuration contradicts the expert view that an improvement in the efficiency η cannot be expected of a nozzle-shaped top disc, as is usual in radial fans having backward-curved blades.

In the case of known radial fans having forward-curved blades, the nozzles do not dip into the top disc. In known radial rotors having forward-curved blades, the static pressure difference at a gap between the nozzle-shaped inlet and the top disc is too small in order for said difference to apply the main flow, moving axially along the longitudinal axis X-X through the inlet, to the top disc by an impulse supply from the split flow also passing laterally through the gap into the suction opening in the top disc. Furthermore, as a result, the blades are subjected to an axial onflow near the top disc, as a result of which the flow is separated at the blade inlet edges.

However, instead, these disadvantages can be avoided in the disclosure by the gap 22 formed by dipping the nozzle-shaped inlet 21 into the suction opening 4 in the top disc 1 over the dipping length L_E. In this respect, the dipping length L_E of the gap 22 can be within a range of 0.5% to 5.0%, preferably within a range of 1.0% to 3.0% of the external diameter D_a,DS of the top disc 1 and a gap width S_W of the gap 22 can be within a range of 0.5% to 5.0%, preferably 1.0% to 3.0% of the external diameter D_a,DS of the top disc 1.

It has proved to be extremely favourable for the formation of the flow downstream of the inlet 21 if an inner radius R_i,S on the leading edge 2a of the blades 2 (see FIG. 2) in the vicinity of the top disc 1 is greater than or equal to the inner radius R_i,DS of the suction opening 4 of the top disc 1 (see FIG. 3).

Furthermore, as shown in FIG. 3, it can preferably be provided that a ratio V_B of a width B of the housing 7 at the air inlet opening 7b thereof into the air guidance channel 7a to an impeller outlet width b₂ of the radial impeller 10 at the radially outer air outlet side 5b of the flow channel 5 has a value within the region of 1.0≦V_B≦1.4. As a result, while avoiding losses of the total pressure Δp_t, the conversion of dynamic pressure into static pressure is promoted. The secondary flow in the housing 7 is positively influenced due to a configuration of this type of the housing 7 with a small increase in width according to the stated ratio V_B, which, contrary to expert opinion, leads to a significant increase in the efficiency η. In this respect, it is particularly advantageous if the impeller outlet width b₂ at the radially outer air outlet side 5b of the flow channel 5 assumes a value which is at most 70% of the overall width b_ges of the radial impeller 10.

Finally, from the point of view of a high efficiency η, it is also advantageous if the housing 7 has an air guidance channel 7a, wound spirally around the radial impeller 10, not with a rectangular, but with an oval, preferably elliptic cross section which increases constantly from the side of the radial impeller 10. In an elliptic cross section of this type, the ratio of the large to small semi-axis of the ellipse can preferably be within a range of 1.2 to 3.0, it being possible for the large semi-axis to be oriented in a different manner, for example preferably vertically or horizontally.

The disclosure is not restricted to the illustrated and described embodiments, but also includes all configurations which have the same effect within the meaning of the disclosure. Thus, a person skilled in the art can also provide expedient additional technical measures without thereby departing from the scope of the disclosure. For example, it can advantageously be provided that an outlet diameter D_a,S of the blades 2 on the top disc 1 is less than or equal to the external diameter D_a,DS of the top disc 1. It can also be provided that this outlet diameter D_a,S is less than or equal to the external diameter D_a,BS of the bottom disc 3.

It has already been stated that the top guiding surface 8 and/or the bottom guiding surface 9 of the flow channel 5 can be curved in a constant manner. The same preferably also applies to the respective pressure sides 2c and suction sides 2d of the blades 2, the above wordings “ . . . at least in part” and “curved in a concave manner at least in regions” (or “curved in a convex manner . . . ”) meaning that the respective curvatures can also comprise straight portions, particularly at the ends thereof.

If the radial impeller 10 is arranged in the housing 7 coaxially with an electric drive motor 6, as shown in FIGS. 3 and 5, an outer contour 6a of the drive motor 6 can engage positively in a motor receiving opening 3a (most clearly visible in FIG. 2) of the bottom disc 3 or alternatively it can also be covered by a full-surface bottom disc 3, the bottom disc 3, preferably together with the contour 6a of the motor 6 received in its motor receiving opening 3a, having a dome-shaped form. When a motor 6 is used in the impeller region, burbling from the motor contour and backflow in the downstream region can be prevented by a dome shape of this type of the bottom disc 3. This is evident, for example, by comparing FIGS. 3 and 5 with FIG. 4, FIG. 4 showing the undesirable formation of vortices W in a wake space between motor 6 and bottom disc 3.

A twin-flow configuration of the radial impeller 10 according to the disclosure is also possible, without departing from the scope of the disclosure.

Furthermore, the disclosure can also be defined by any other combination of particular features of all individual features disclosed overall. This means that in principle, practically every individual feature can be omitted or replaced by at least one individual feature disclosed elsewhere in the application.

The foregoing description of the embodiments has been provided for purposes of illustration and description. It is not intended to be exhaustive or to limit the disclosure. Individual elements or features of a particular embodiment are generally not limited to that particular embodiment, but, where applicable, are interchangeable and can be used in a selected embodiment, even if not specifically shown or described. The same may also be varied in many ways. Such variations are not to be regarded as a departure from the disclosure, and all such modifications are intended to be included within the scope of the disclosure.

LIST OF REFERENCE SIGNS

• 1 top disc of 10
• 2 blades of 10 between 1 and 3
• 2a leading edge of 2
• 2b trailing edge of 2
• 2c pressure side of 2
• 2d suction side of 2
• 3 bottom disc of 10
• 3a motor receiving opening for 6 in 3
• 4 suction opening of 1
• 5 flow channel in 10 between 2/2 and between 8/9
• 5a air inlet side of 5
• 5b air outlet side of 5
• 6 drive motor for 10
• 6a outer contour of 6
• 7 housing of 20
• 7a air guidance channel of 7
• 7b air inlet opening of 7 into 7a
• 8 top guiding surface of 5
• 9 bottom guiding surface of 5
• 10 radial impeller
• 20 fan unit with 10
• 21 inlet of 7
• 22 gap between 21 and 1
• A shortest distance between 1 and 3
• B width of 7 at 7b
• b₂ impeller outlet width of 1
• b_DS width of 1
• b_ges overall width of 10
• D_a,BS external diameter of 3
• D_a,DS external diameter of 1
• D_a,S outlet diameter of 2 at 1
• D_i,DS internal diameter of 4 in 1
• d_S profile thickness of 2
• LR running direction of 10
• L_E dipping length of 21 into 1, length of 22
• L_S profile overall length of 2
• R_i,S inner radius of 2 at 2a
• R_i,DS inner radius of 4 in 1
• S flow direction in 5 from 5a to 5b
• T₁ tangent on 8 at 4
• T₂ tangent on 8 at 5a
• W vortex between 3 and 6 (FIG. 4)
• X-X longitudinal axis of 10, 20
• α_DS angle between T₁ and T₂
• α_DS1 angle between T₁ and X-X
• α_DS2 angle between T₂ and X-X

Claims

1. A radial impeller having an axial air inlet and an air outlet over an impeller circumference, preferably for use in a spiral housing, the radial impeller comprising:

a bottom disc which has an external diameter,

a top disc which is arranged concentrically to the bottom disc at an axial distance therefrom and has a circular suction opening with an internal diameter for the axial air inlet, and

a plurality of forward-curved, profiled blades arranged between the bottom disc and the top disc, a flow channel having a radially inner air inlet side and a radially outer air outlet side being formed between two adjacent blades in each case, which channel is curved convexly, viewed in the running direction of the radial impeller, wherein

the external diameter, of the bottom disc, based on the internal diameter of the suction opening, is at least 20% greater than the internal diameter of the suction opening, in that the top disc forms a top guiding surface for the flow channel, an angle which opens towards the inlet opening and is formed between a tangent on the top guiding surface at the inlet to the suction opening of the top disc and a tangent on the top guiding surface at the outlet from the flow channel on the radially outer air outlet side thereof, being at least 30°.

2. The radial impeller according to claim 1, wherein the external diameter of the bottom disc, based on the internal diameter of the suction opening, is at least 50%, but at most 90% greater than the internal diameter of the top disc.

3. The radial impeller according to claim 1, wherein in the profiled blades, a profile thickness ratio of maximum profile thickness to profile overall length is at least 0.15, in particular at least 0.2 and more preferably at least 0.25, the profile thickness ratio being at most 0.5 and in particular 0.4, more preferably 0.35.

4. The radial impeller according to claim 1, wherein an inner radius on the leading edge of the blades in the vicinity of the top disc is greater than or equal to the inner radius of the suction opening in the top disc.

5. The radial impeller according to claim 1, wherein an outlet diameter of the blades on the top disc is less than or equal to the external diameter of the top disc and/or is less than or equal to the external diameter of the bottom disc.

6. The radial impeller according to claim 1, wherein a number of blades is at least 19 and at most 54 and is preferably within a range of 22 to 46.

7. The radial impeller according to claim 1, wherein the top disc, the blades and the bottom disc are configured as a composite body consisting of two parts, in particular of two plastics injection-moulded parts which are integrally bonded.

8. The radial impeller according to claim 1, wherein the bottom disc forms a bottom guiding surface in the flow channel.

9. The radial impeller according to claim 1, wherein the top guiding surface and/or the bottom guiding surface of the flow channel have a constant curvature.

10. The radial impeller according to claim 1, wherein in each case the leading edges of the blades on the radially inner air inlet side and/or the trailing edges of the blades on the radially outer air outlet side are rounded.

11. The radial impeller according to claim 1, wherein the cross section of the flow channel tapers from the radially inner air inlet side to the radially outer air outlet side, in particular a shortest distance between the bottom disc and the top disc decreasing in the flow direction.

12. The radial impeller according to claim 1, wherein an impeller outlet width on the radially outer air outlet side of the flow channel has a value which is at most 70% of an overall width of the radial impeller.

13. The radial impeller according to claim 1, wherein the overall width thereof is within a range of 25% to 70%, based on the external diameter of the bottom disc.

14. The radial impeller according to claim 1, wherein the angle which opens out towards the inlet opening and is formed between a tangent on the top guiding surface at the inlet to the suction opening of the top disc and a tangent on the top guiding surface at the outlet from the flow channel on its radially outer air outlet side is at most 90°, preferably at most 75°.

15. A fan unit, in particular a drum fan, having a radial impeller according to claim 1, wherein the radial impeller is arranged in a housing, in particular in a spiral housing.

16. The fan unit according to claim 15, wherein an inlet of the housing for the radial impeller is nozzle-shaped, the inlet of the housing in particular dipping into the suction opening in the top disc.

17. The fan unit according to claim 15, wherein a ratio of a width of the housing at the air inlet opening thereof into an air guiding channel to an impeller outlet width of the radial impeller at the radially outer air outlet side of the flow channel has a value within a range of 1.0≦VB≦1.4.

18. The fan unit according to claim 15, wherein the radial impeller is arranged in the housing coaxially with an electric drive motor, an outer contour of the drive motor engaging positively in a motor receiving opening in the bottom disc or being covered by a full-surface bottom disc, and the bottom disc, preferably together with the contour of the motor received in the motor receiving opening thereof, having a dome-shaped formation.

19. The fan unit according to claim 15, wherein the housing has an air guidance channel, wound spirally around the radial impeller, with an oval cross section which increases constantly from the side of the radial impeller.