Turbulent Flow Measurements near the Discharge Port of a Screw Compressor
Mean flow and turbulence characteristics have been measured within the male and female rotors close to the discharge port of a double screw compressor at different radial positions, two axial positions from the exit port, Hp, and two radial planes, αp. Cycle-resolved axial and tangential mean flow measurements and their corresponding turbulent velocity fluctuations were made over a time window of 1° using a laser Doppler velocimetry, LDV, system. Measurements were performed through two transparent windows near the inlet of the discharge port inside the male and female working chambers. The results revealed a highly complex 3-D flow within the male and female working chambers, in particular, near the discharge port with two distinct flow zones 1 and 2 before and after the opening of the port, respectively. The flow in zone 1 was controlled by the rotor motion while in zone 2 was greatly influenced by the discharge process. In zone 2, both components of mean velocities were subjected to a sudden increase in velocity forming strong axial and tangential jet flows due to rapid change in pressure across the port as the flow is exposed into the discharge port. It was found that the flow structures have been affected considerably by the position of the discharge port, radial planes and radial positions. Axial and tangential RMS velocity distributions within both rotors were found to be relatively high and less affected by the flow changes of zones 1 and 2 with almost uniform distribution. The measured magnitudes of axial and tangential RMS velocities suggest it would be reasonable to assume the local turbulence to be isotropic for the modelling purposes. To authors’ knowledge, the results are unique, original and in great details not only to describe the flow structure, but also, they can be used in CFD codes to establish a reliable model of the flow and pressure distribution within twin screw machines.
KeywordsScrew optical compressor Laser Doppler Velocimetry Cycle-resolved averaging Turbulent flows Mean and RMS velocities
Screw compressors could be characterised with their simple arrangement and compact design, high rotational speed and high efficiency over a wide range of speeds and pressure differences, which make them suitable for many industrial applications. Their use nowadays is widespread since they have replaced the traditional reciprocating compressor. Screw compressors are used commonly in a large range of applications such as air compression, refrigerant compression, fuel cell and turbo charging for the automotive industry and many others like building industry, food process, pharmaceutical, chemical industry, metallurgical industry and pneumatic transport. At present, they constitute a large percentage of all positive displacement compressors that are sold and currently in operation. The main reason for this success is the development of new rotor profiles and advanced machine tool, which can manufacture the rather complex rotor shape, to tolerances of the order of 3 μm at an acceptable cost . Improvements in screw compressors are continually sought-after, in order to increase their performance, reduce their energy consumption, noise generation and manufacturing costs. A screw compressor consists of two rotors, male and the female, contained in a casing with no valves and their meshing lobes form a series of working chambers within which compression takes place as described in [1, 2, 3]. As the rotors turn, air is admitted through the space between the rotor lobes and the suction port. Further rotation of the rotors lead to a cut-off of the suction port and the trapped air is pushed forward axially and circumferentially towards the discharge port by the action of the screw rotors; during this period the trapped volume in each passage is reduced and its pressure is increased. This process continues until the working volume between the rotors is exposed to the discharge port allowing gas at high pressure to flow out [2, 3]. Flow leakages through clearances and tolerances take place across several lines like the contact between the two rotors, the sealing between the rotors’ tip and the casing, and the clearance between the rotor end face at the discharge end plate.
The entire flow process and, in particular, the leakages play an important role in the compressor performance. It is thus essential to have a good understanding of the gas flow motion in the compressor by quantifying the velocity field in the compressor elements, suction, discharge and working chambers and, especially, through the clearance gaps (if possible) so as to characterize the whole sequence of processes that occur within the compressor. The total number of papers published on the flow characteristics within twin-screw compressors is rather small compared to published work on other machines such as turbochargers and pumps. Moreover, the investigations of [4, 5, 6, 7] seem to be the only reported experimental works on the flow behaviour in screw compressors or screw superchargers who have reported the axial mean and RMS velocities within the working chamber of the male rotor only at one radial plane of αp = 27° using LDV system. They have reported that the chamber-to-chamber (cyclic) flow variations were negligible and the flow velocities of all five working chambers were very similar. They also reported that the axial flow distribution across the working chamber of male rotor were separated in three zones 1, 2 and 3 controlled by the rotor motion, by the discharge process and by the flow leakage near the tip, respectively. The present research work is the continuation of the works of [4, 5, 6, 7] and it is intended to complete the mapping of the flow on the same screw compressor for both male and female working chambers. The velocity field measurements are also expanded to include both axial and tangential mean and RMS velocity characteristics at different radial planes (αp = 17° and 37°) that were reported by [4, 5, 6, 7]. Thus, to the authors’ knowledge, the presented results in this report are unique contribution and not available in the open literature. However, there are many experiments focusing on the performance of the compressors and CFD calculations; for example, [8, 9, 10] describe some performance test measurements in the same compressor used for this experiment where properties of other parameters in the compressor, such as the suction and discharge pressures and temperatures have been measured with standard laboratory-type instruments and compared with predicted values from an existing CFD model, which allowed the validation and further development of the CFD [11, 12, 13] package. In addition, [14, 15, 16] reported theoretical and experimental studies on screw compressors and made transient thermal analysis, numerical simulation and experimental verification. Presented a procedure to determine empirical constants in a ‘training process’ which optimises multiple constants using a range of test cases. [14, 16] presented models that after proper calibration gave good agreement between simulated and measured results, while  showed that thermal deformation is accounted for by complex 3D FEA or CFD calculations when the nominal design clearances is corrected with the measured correction factors.
The material presented in this paper is part of a long-term research project attempting to measure the fluid mean velocity distribution and the corresponding turbulence fluctuations at various cross-sections within the working chambers of the male and female rotors to characterise the mean and turbulent flow development through the compressor at different phase angles. The overall aim is to reveal how major features of the fluid flow within the machine are affected by the rotor geometry and operating conditions. Flow in screw compressors is complex, three-dimensional and strongly time-dependent as already indicated by the results of [4, 5, 6, 7] and are similar in flow complexity to the in-cylinder flows of IC engines [17, 18], centrifugal pumps  or in turbocharge turbines [20, 21, 22] and mixing reactors . This implies that the measuring instrumentation must be robust to withstand the unsteady aerodynamic forces, have high spatial and temporal resolutions and, most important, must not disturb the flow. Only the optical diagnostics like LDV and PIV  can fulfil these requirements, in particular, resolving the turbulent flow characteristics with LDV as successfully demonstrated by previous research in similar complex flows of [17, 18, 19, 20, 21, 22, 23].
The preferred method of research is characterisation of the fluid mean velocity and turbulence fluctuations at a range of pre-selected measurement points using a dual beam Laser-Doppler Velocimetry (LDV). Angle-resolved of axial and tangential mean and RMS velocities have been obtained inside the working chambers of the male and female rotors. Measurements were performed in the interlobe region near the discharge port by passing the laser beam through purposed built transparent windows, made of Plexiglas (Perspex), that were installed on a grove machined on the compressor casing; more details will be given in the following section. The results will be presented and discussed in the subsequent section and the reports will end with a summary of the main findings.
2 Flow Configuration and Instrumentations
The desired LDV measurements within the compressor under oil flooded (for cooling purposes) conditions were not possible due to sever and continuous fouling of the optical window as explained in [5, 6, 7]. Thus, in order to be able to measure the velocity field, it was decided to run the compressor with no oil injection at a speed of 1000 rpm with outlet and inlet pressure ratio 1:1 and a discharge temperature of 55 ± 3 °C. Despite this unusual working condition, the internal compression ensures a flow structure similar to the one occurring in normal operating conditions. The flow regime was identified by the Reynolds number defined as Re = ρVpDp/μ where ρ and μ are air density and dynamic viscosity, respectively, where Vp is the male rotor pitch axial velocity defined as the axial speed at which a single working chamber travels from the suction to the discharge port of the compressor and is given by Vp = ωDptanβ /2 where ω = 2πN/60, Dp is the male rotor pitch diameter and, β is the pitch angle. Vp was calculated to be 3.95 m/s for a rotor speed, N, of 1000 rpm, pitch diameter of 81.8 mm and pitch angle of 42.73°. The average Reynolds number was found to be 17,700 and therefore the flow can be considered to be turbulent; this is based on consideration of an axial flow through the working chamber to be similar to that of a typical helical duct flow, and that all angled-resolved (over a 1°) velocity measurements clearly indicated that the flow is turbulent. The volumetric flow rates through the compressor were measured by an orifice plate installed in the exhaust pipe and it was 1.05 m3/min. The speed was obtained with an optical shaft encoder which was fixed at the end of the driving shaft with 3600 train pulses per revolution providing an angular resolution of 0.1 degrees. With this arrangement, the male rotor was rotating clockwise and the female anticlockwise as shown in Fig. 3.
The air flow was seeded by a silicone oil atomiser capable of producing droplet sizes in the range of 1 to 2 μm; a low viscosity silicone oil of 5 cSt was used. A Matlab program was written in which the information of the shaft angular position from the shaft encoder was used to resolve the velocity with respect to the rotors’ angular position, the so-called ‘gated’ measurements. This has been done by collecting the sum of all the instantaneous velocities over a given time-window and then the ensemble mean and RMS values were calculated. This method of gated measurements proved to be efficient since the data was collected continuously over the entire 360° cycle as the rotor turned, and provided ensemble averages for every one degree (∆θ = 1°) time-window. The total measuring time for each point was about 25 min which, apart from a few critical points, gave a minimum number of 700 samples per time-window corresponding to statistical uncertainties of less than 1.7% and 5.5% in the ensemble mean and RMS velocities, respectively, based on a 95% confidence level and velocity fluctuations of the order of 20% of the mean value. The direction of the measured axial velocity component is along the axes of the rotor with positive value towards the discharge port. For tangential velocity measurement, the optical set up provided a positive value in the clockwise direction.
3 Results and Discussion
3.1 Mean Flow Velocities
Figure 6 compares mean axial flow variations inside the working chamber of the male (left column) and Female (right column) with rotors’ angular positions, θ, at a distance Hp = 34 mm from discharge port, along two radial planes αp = 17° and 37°, at different radial locations, Rp, from near the root of the rotors up to their tips, and for a speed of 1000 rpm. The axial flow variations within the working chamber of the male rotor at αp = 27° were presented and described by [4, 5, 6, 7] and only a summary is given here. First, the chamber-to-chamber (cyclic) flow variations were established and the results showed that the axial mean and RMS velocities of all five working chambers were very similar with small differences near the leading edge of the rotor. The axial velocity distribution across the working chamber could be described by dividing the working chamber into three different zones. The first zone was in between the LE and the opening of the discharge port, θ = 0°, where the flow was controlled by the rotor motion with velocity profiles showing a gradual decrease towards θ = 0°; the results also showed that at lower Rp (close to the root of the chamber) the velocities across the chamber were generally lower and became highest at higher Rp, near the tip. The second zone was when the flow was exposed to the discharge port after θ = 0 and was influenced by the discharge process and relative position of the of the port’s exposed-area. The third zone was restricted to the vicinity of the LE near the tip, where the flow is influenced mainly by flow leakage between two adjacent working chambers and is generally characterised by high velocity. All these will be reassessed when discussing the axial and tangential flow at radial planes of αp = 17° and 37°.
The male axial mean flow variation across the working chamber (left column) clearly shows the presence of all these 3 zones for both radial planes of αp = 17° and 37° with an exception that zone 2 is almost absent in the case of αp = 17°, which suggests that the flow is not yet exposed to discharge port at this radial plane due to physical position of the port which is situated in the central part of the interlobe region covering larger radial planes (αp) and at larger Rp; the latter also explains the absence of the zone 2 flow characteristics near the root of the male rotor and lower mean velocity values there. It is also evident from the graphs that velocity values at plane αp = 37° are higher than those at plane αp = 17° for the same reason given above and that the angular axial velocity profiles near the root (lowest Rp) is much narrower than those near the tip due to large variation of the internal profile of the male rotor’s blade with Rp, see Fig. 3. It is interesting to note that the velocity profile near the root looks almost like a parabola (similar to a duct flow with no slip conditions at blade walls), but as Rp increases the velocity values, in general, increases and the profiles spread across the working chamber with the maximum velocity closer to the LE and reducing gradually towards the TE. Another important flow feature for the male rotor is the flow variation in zone 2 after the port opening as is evident in Fig. 6(b) left column. The results show that the velocities continue to decrease up to θ ≈ 5° but by a faster rate than before the opening of the port, and then they increase rapidly towards the TE. The initial decrease indicates the presence of an adverse pressure gradient downstream as the port opening hasn’t exposed to this radial plane yet, while the flow at θ ≈ 5° starts to sense the exposed area of the port with large pressure difference across the port and therefore rapid increase in velocity. In Zone 3, the axial flow velocities are highest in the vicinity of the LE near the rotor tip (for Rp > 60 mm). This is mainly due to the flow leakage between two adjacent working chambers. Since the rotors have a helical shape then the leakage flow on the male rotor as it enters into the next chamber would have strong axial and tangential components towards the port and tangential directions adding to the axial and angular momentums (or increase in their velocities) within the chamber with its highest impact to be close to LE and near the tip.
The mean axial flow characteristics within the female rotor working chamber (right column) of Fig. 6 show clearly the presence of only two distinct zones 1 (from LE to θ = 0°) and 2 (from θ = 0° to TE) with no flow leakage (zone 3) detected on the female rotor at both radial planes compared to the male rotor where strong leakage flow was detected. The reason for this may well be due to the different shape of the rotors at the contact point with the compressor casing where with female rotors the blade has almost straight TE and LE profiles normal to the casing (see Fig. 3), while with the male rotors has a strong convex LE and TE profiles forming a convergent divergent flow passage across the contact point, which creates a relatively strong pressure difference across the contact point as the rotor turns and hence promoting the leakage flow. Within the zone 1, the flow is mainly controlled by the rotor motion, similar to the male rotor, and shows generally uniform angular axial velocity profiles across the working chamber, in particular at αp = 37°; this is in contrast with the male rotor, which may be due to more uniform chamber cross-section area with female rotor than that of male rotor. In the second zone (from θ = 0° to TE), the flow is controlled by the discharge process and the results at αp = 37° clearly shows the flow is associated with a sudden and rapid flow velocity increase as it is exposed to the opening of the port at θ ≈ 3° till very close to TE of the rotor (θ ≈ 10°) where the flow velocity starts to drop near the TE wall as it would be expected. Again, the second flow zone is absent at αp = 17° because the flow is not yet exposed to the discharge port at this radial plane as explained above. It is interesting to note that the flow velocities at αp = 37° are almost twice as large as those at αp = 17° with values similar to that of Vp. Also, unlike the male rotor, the angular axial velocity profiles at different Rp locations, from the root to near the tip, span over a similar range of θ due to again the much smaller variation of the internal profile of the two adjacent female rotor’s blades with Rp compared to that of male rotor, see Fig. 3.
Closer to discharge port at Hp = 22 mm, Fig. 9, tangential mean flow variations within the male and female working chambers show similar, but more pronounced, flow structures to those observed at Hp = 34 mm with velocity magnitude slightly larger particularly in zone 1. Flow structures of zones 1 and 2 are clearly evidence at both radial planes before and after the opening of the discharge port, θ = 0°, in particular, in zone 2 at αp = 17° where the influence of the discharge port opening is evidence in formation of an anti-clockwise tanjential flow jet with maximum velocities of values up to -2Vp at all radial positions, Rp.
Figures 6, 7, 8 and 9 clearly show the presence of a complex 3-D mean flow within the male and female near the discharge port and that the opening of the discharge port influence both axial and tangential flow within both rotors as they approach the exposure area of the exit port. To summarise, it can be concluded that, the results obtained for all axial and tangential components, within both male and female rotors’ chambers, follow three distinct flow zones 1, 2 and 3 as described above, and are influenced by the rotors’ motions, opening of the discharge port and the chamber-to-chamber flow leakage, respectively. For the latter flow zone 3, no clear flow leakage was detected for the female rotors.
3.2 Turbulent Flow Velocity Fluctuations
Figure 11 presents the axial RMS velocity distributions within the male and female working chambers for radial planes of αp = 17° and 37° at Hp = 22 mm closer to the discharge port where the whole flow is more exposed to the discharge port. The results show, unlike at Hp = 22 mm, the axial RMS velocities are less uniform with θ and higher fluctuation with Rp than those observed in Fig. 10 further away from the port. This is expected as the larger mean flow variations were observed at this axial location as the flow is more exposed to discharge port. Despite these, the overall average RMS value at this axial station remain similar to that of Hp = 34 mm with a value of around 0.3–0.35Vp. It should also be noted that although the fluctuation of RMS velocity with Rp is larger at this location, still there is no clear trend to indicate that the changes can be correlated to Rp. Again, the largest RMS velocity values are with the male rotor with the similar magnitude to those detected at Hp = 34 mm near the LE and the rotor’s tip within the flow of zone 3 associated with the tip flow leakage.
Again, there is no particular trend between the RMS angular velocity profiles within the male rotor (left column) working chamber with the radial positions, Rp, with overall values at around 1.75 m/s (0.43Vp). However, with female rotor (right column), there is a clear correlation between the RMS variation and the radial positions, Rp, so that the RMS level is lowest at the root of the rotor and increases with Rp to become largest near the tip of the rotor; the minimum and maximum values are of order of 1 m/s (0.25Vp) and 1.75 m/s (0.43Vp), respectively; similar range that was observed for axial components. Considering the overall RMS results of axial and tangential components for both rotors, it would be reasonable to assume the local turbulence to be isotropy for the modelling purposes.
The mean axial and tangential velocity distributions within the male and female working chambers confirmed the presence of two distinct flow zones 1 and 2 before and after the opening of the discharge port, θ = 0°, controlled mainly by the rotor motion and the flow exposure into the discharge port, respectively. A third zone was also identified with male rotor only near the tip region, where the chamber flow is influenced greatly by the leakages through the gaps between the rotors and the casing.
In zone 1 at both radial planes, αp, the axial mean flow within the male chamber was associated with an increase of velocity from the leading edge (LE) to a maximum value around the core region and then decreased towards port opening (∆θ = 0°), while the tangential mean flow showed an opposite flow pattern with a minimum value around the core region.
The largest changes in mean flow velocities were observed in zone 2, where a sudden increase in mean axial and tangential velocity components were obtained in the form of strong jet-like flows within both male and female working chambers due to opening of the discharge port and the exposure of the high-pressure flow into the discharge cavity; these flow features were more pronounced closer to the discharge port at Hp = 22 mm.
RMS velocity distributions within the male and female working chambers were found to be uniform across the chamber with small differences at different Rp and θ; the differences were more pronounced closer to the discharge port at Hp = 22 mm. Despite that, the overall average RMS value at both Hp remained similar with values around 0.3Vp and 0.35Vp for axial and tangential flows, respectively.
Overall, the mean and turbulent results clearly showed the presence of a 3-D mean flow structure within the male and female chambers that became more complex as it moved closer to the discharge port. The similarity of the measured turbulent levels for both components and rotors suggested that it would be reasonable to assume the local turbulence to be isotropy for the modelling purposes.
The new measured data would be a unique contribution to literature, and have been obtained in great details to be used for the validation of CFD codes to establish a reliable model for accurate prediction of flow and pressure distribution within twin screw machines, which can then be used as a tool to further improve the design of screw compressors and expanders.
Financial support from EPSRC is gratefully acknowledged. The authors would like to thank Tom Fleming, Michael Smith, Jim Ford and Grant Clow for their valuable technical support during the course of this work.
- 1.Stosic, N.: On gearing of helical screw compressor rotors. Proc IMechE. 213(C), 587–594 (1998)Google Scholar
- 3.Guerrato, D.: Cycle-Resolved flow Characterisitcs Within a Screw Compressor, Ph.D. Thesis, City, University of London, 2012Google Scholar
- 4.Nouri, J.M., Guerrato, D., Stosic N., Kovacevic, A.: Cycle-resolved velocity measurements within a screw compressor. 18th Int. Compressor Eng. Conf., Purdue, USA, 17–20 July 2006Google Scholar
- 5.Guerrato, D., Nouri, J.M., Stosic, N., Arcoumanis, C.: Axial flow and pressure characteristics within a double screw compressor. 3rd Int. Conf. Optical and Laser Diagnostics, ICOLD, London, 23–25 May 2007Google Scholar
- 6.Nouri, J.M., Guerrato, D., Stosic, N., Arcoumanis, C.: Axial flow characteristics within a screw compressor. ASHRAE Research Journal. 14(2), 259–274 (2008)Google Scholar
- 7.Guerrato, D., Nouri, J.M., Stosic, N., Arcoumanis, C., Smith, I.K.: Flow measurements in the discharge port of a screw compressor. Proceedings of the IMechE, Part E: Journal of Process Mechanical Engineering. 222(E4), 201–210 (2008)Google Scholar
- 8.Kovacevic, A., Stosic, N., Smith, I.K.: The CFD Analysis of a Screw Compressor Suction. Int. Compressor Eng. Conf., Purdue, USA 909 (2000)Google Scholar
- 9.Kovacevic, A., Stosic, N., Smith, I.K.: CFD analysis of screw compressor performance. In: Tourlidakis, A., Yates, M.K., Elder, R.L. (eds.) Advances of CFD in Fluid Machinery Design. Professional Engineering Publishing, London (2002)Google Scholar
- 10.Stosic, N., Mujic, E., Kovacevic, A. Smith, I.K.: The Influence of discharge ports on rotor contact in screw compressor. 18th Int. Compressor Eng. Conf., Purdue, USA 17–20 July 2006Google Scholar
- 15.Sauls J, Powell G, Weathers B.: Transient thermal analysis of screw compressors part I - development of thermal model. VDI Berichte, 19–29 (2006)Google Scholar
- 16.Hsieh, S.H., Shih, Y.C., Hsieh, W.H., Lin, F.Y., Tsai, M.J.: Performance analysis of screw compressors – numerical simulation and experimental verification. Proc. IMechE part C, J. mechanical engineering. Science. 226, 968–980 (2012)Google Scholar
- 17.Kampanis, N., Arcoumanis, C., Kato, R., Kometani, S.: Flow, combustion and emissions in a five-valve research gasoline engine. SAE Paper 2001–013557, Ditroit, USA, 2001Google Scholar
- 18.Yan, Y., Gashi, S., Nouri, J.M., Lockett, R.L., Arcoumanis, C.: Investigation of Spray Characteristics in a Spray Guided DISI Engine Using PLIF and LDV, 2nd Int. Conference on Optical Diagnostics, ICOLD, London, September 2005Google Scholar
- 19.Liu, C.H., Nouri, J.M., Vafidis, C., Whitelaw, J.W.: Experimental Study of Flow in a Centrifugal Pump, 5th International Symposium on Applications of Laser Techniques to Fluid Mechanics, Lisbon, July 1990Google Scholar
- 21.Arcoumanis, C., Martinez-Botas, R., Nouri, J.M., Su, C.C.: Inlet and exit flow characteristics of mixed flow turbines. ASME, Int. gas Turbine and Aeroengine Congress and Exhibition, Paper 98-GT-495, Stockholm, June 1998Google Scholar
- 22.Zaidi, S.H., Elder, R.L.: Investigation of flow in a radial turbine using laser anemometry. ASME, Int. Gas Turbine and Aeroengine Congress and Exhibition, Cincinnati, paper 93-GT-55, Ohio, May 1993Google Scholar
Open Access This article is distributed under the terms of the Creative Commons Attribution 4.0 International License (http://creativecommons.org/licenses/by/4.0/), which permits unrestricted use, distribution, and reproduction in any medium, provided you give appropriate credit to the original author(s) and the source, provide a link to the Creative Commons license, and indicate if changes were made.