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Twin screw oil-free wet compressor for compression–absorption cycle
International Journal of Refrigeration 29 (2006) 556–565 www.elsevier.com/locate/ijrefrig
Twin screw oil-free wet compressor for compression–absorption cycle C.A. Infante Ferreira1,*, C. Zamfirescu, D. Zaytsev2 Mechanical, Maritime and Materials Engineering, Delft University of Technology, Mekelweg 2, 2628 CD Delft, The Netherlands Received 10 February 2005; received in revised form 15 September 2005; accepted 10 October 2005 Available online 20 December 2005
Abstract The performance of a twin screw compressor operating under wet (two-phase) compression conditions in an ammonia–water compression absorption heat pump cycle is investigated both theoretically and experimentally. The paper reports on the influence of the location of liquid intake or, depending what applies, injection angle and mass flow rate of the injected liquid on compressor performance. Labyrinth seals separate the oil-free process side from oil lubricated bearing housing. Labyrinth seals leakage is modelled and its impact on performance is theoretically and experimentally investigated. The need for liquid injection from the discharge side to obtain acceptable performance is discussed based on experimental results. q 2005 Elsevier Ltd and IIR. All rights reserved. Keywords: Absorption system; Ammonia-water; Screw compressor; Oil-free compressor; Research; Experiment; Performance
Compresseur a` double vis secs dans un cycle a` compression a` absorption Mots cle´s : Syste`me a` absorption ; Ammoniac-eau ; Compresseur a` vis ; Compresseur sec ; Recherche ; Expe´rimentaion ; Performance
1. Introduction Compression–absorption heat pump cycles have the potential to significantly improve the performance of heat pump cycles based on the vapor compression principle. Itard [1,2], Hulte´n and Berntsson [3] and more recently Zaytsev
* Corresponding author. Tel.: C31 152784894; fax: C31 152782460. E-mail address: [email protected] (C.A.I. Ferreira). 1 Member of IIR Commission B1. 2 Present address: Grasso GmbH RT, R&D Screw Compressors, Holzhauser Strasse 165, 13509 Berlin, Germany.
0140-7007/$35.00 q 2005 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2005.10.006
[4] investigated this potential. For high temperature industrial heat pumps, a gain of about 20% in comparison with ‘conventional’ vapor compressor heat pumps can be attained. Two different ways of implementing compressor– absorption heat pump cycles have been investigated in the past: the wet compression and the dry compression types. The differences between the two solutions have extensively been discussed by Itard [1,2]. The dry compression implementation raises two problems: (i) the irreversibility losses associated with vapor superheating play a major role in degrading the overall performance of the cycle; (ii) dry compressors require oil lubrication, with the implication of oil contamination of the liquid refrigerant [5] and additional degradation of the performance. Since the liquid refrigerant
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Nomenclature Aef ALab CD C0 D h j l m m_ n N p Q T u v
leakage flow area (m2) labyrinth leakage flow area (m2) empirical discharge flow coefficient empirical leakage paths flow coefficient labyrinth blade diameter (m) enthalpy (J kgK1) number of inflow paths length (m) mass (kg) mass flow (kg sK1) number of outflow paths number of labyrinth blades pressure (Pa) heat (J) temperature (K) internal energy (J kgK1) specific volume (m3 kgK1)
and the oil have similar densities, very expensive separation devices must be included in the heat pump cycle. Oil-free compressors operating under wet conditions avoid superheating. The liquid refrigerant is used as lubricant and the oil—still needed for lubrication of the bearings—is separated from the compression process. The performance advantage of the wet compression in comparison with the dry compression varies from 2.5 to 13.0% for the application cases studied by Itard [1]. However, the lack of information related to two-phase compression processes, and the use of liquid refrigerant itself as a lubricant, makes the development of a suitable ‘wet compressor’ difficult. Zaytsev [4] gives an overview of compression–absorption cycle investigations that have been carried out in recent years. Most investigations have been dedicated to the dry compression type. Very recently, Satapathy et al. [6] and Risberg et al. [7] reported on dry compression systems. Malewski [8], Bergmann and Hivessy [9], Itard [1] and Zaytsev [4] have reported experimental studies on wet compression absorption cycles. Zaytsev [4] reviewed the different compressor types available in the market for their applicability as compressor for wet compression absorption cycles. Considering the specific requirements of these cycles, the screw compressors appeared to be the most suitable choice. They are tolerant to liquid carry over and have rather high efficiency. Zaytsev modified an existing twin-screw compressor available on the market in order to meet the requirements imposed by the wet compression absorption application, implemented it in an experimental compression absorption heat pump and collected several sets of experimental data with it. The isentropic efficiency of this first prototype was rather
V W x z
volume of cavity (m3) velocity (m sK1) ammonia concentration number of rotor’s lobes
Greek letters d clearance (m) r density (kg mK3) 4 rotation angle Subscripts 0 overall 1 male rotor 2 female rotor ef effective inj injection port Lab labyrinth UPR uniform property region
low because the liquid refrigerant was injected into the suction cavity where it flashed into vapor, reducing the volume of displaced vapor significantly. Zaytsev [4] suggested that the compressor efficiency might be improved if the low pressure liquid is separated from the vapor and injected, together with the labyrinth leakage flow (needed to allow for oil-free operation), into the cavity at the start phase of compression. This paper reports the experimental and theoretical investigations carried out to more accurately understand and model the physical processes related to wet compression. Aim is to increase the isentropic efficiency of the compressor to a level that makes wet compression– absorption heat pumps a competitive technical option. Further details of this investigation can be found in Zamfirescu et al. [10]. 2. Experimental set-up The experimental data collected by Zaytsev [4] allowed to re-design the rotors taking thermal expansion into account. The newly built compressor prototype has reduced clearances (at rotor tips, discharge end face, and labyrinth seal), and only one injection port. The experimental ammonia–water compression–absorption set-up previously used by Zaytsev [4] has been used after slight modifications, see Fig. 1. It allows for experimental determination of the compressor performance indicators including indicated diagrams. Also data related to the labyrinth leakage flow have been measured. For data processing purpose the heat pump was split in three functional parts delimited by three states defined as follows: (i) state A is a combination of states A1 and A2, i.e. compressor suction and liquid injection from low pressure
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Fig. 1. Experimental set-up with sample experimental data.
side (at flow meter ‘FT4’); (ii) state B represents the absorber inlet; the properties in state B are a combination of the properties in the liquid line (at mass flow meter ‘mFT5’) and in the vapor line; (iii) state C is located at the absorber outlet. Six Kistler high-frequency piezo-electric pressure transducers in combination with an angle transducer have been used for measurement of the compressor pV-diagram. The first pressure transducer is situated 98 before the compression start point, while the last sensor is placed 258 after the discharge start point. The compressor shaft power is obtained from the driving torque (Koyowa torque transducer) and shaft speed measurements. Further details of the experimental set-up, estimated accuracy of the measurements and resulting uncertainties for relevant parameters can be found in [4]. Table 2 includes a column where the maximum experimental uncertainties of the considered performance indicators are given. The large uncertainty for the efficiencies originates mainly from the calculation of the ammonia concentration.
3. Thermodynamic model Two options of modeling the wet compression of ammonia–water solution were analyzed: a pu-model and a pT-model. Due to the high nonlinearity of the equation of state, the pu-model requires a time consuming iterative procedure. Although the governing equations of the pTmodel are more complicated than those of the pu-model, with the existing ammonia–water subroutines based on Ziegler–Trepp [11] the computation time needed to integrate the pT-model governing equations is shorter. The pressure and temperature are the input variables of the mixture thermodynamic state model; therefore, there is no need of additional iterations to calculate the rest of the mixture state parameters. The model assumes that liquid and vapor are at thermodynamic equilibrium at any moment and that the thermodynamic state of the solution is given by a single set of governing equations (homogeneous model). The
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governing equations for the control volume (cavity) under study are (see [4,12] for further details): The mass conservation equation: dp 1 Z vv d41
vp T;x
C
"
n X
v m
kZ1
dmout d41
! j X dmin K d41 k k kZ1
1 dV vv dT vv dx0 : K K m d41 vT p;x d41 vx0 p;T d41
(1)
The ammonia mass conservation equation: " # j j X dx0 1 X dmin dmin Z x Kx0 : m kZ1 0in;k d41 k d41 d41 k kZ1
(2)
The energy conservation equation: dT Z d41 T
vv
v vT p;x m
n P dmout kZ1
d41
k
K
kZ1 vv vh vp T;x vT
dQ d41
C
C
j P dmin d41
k
p;x C T
C m1
vv 2
vT p;x
vh in ðhin;k KhÞ dm d41 k Km vx0 p;T kZ1 vh vv 2 mT vT m vT p;x C vv p;x j P
dx0 dV vv d41 K vx0 p;T d41
dx0 d41
:
ð3Þ
vp T;x
A numerical integration of the governing equations is used to calculate the three thermodynamic parameters of the mixture: pressure, temperature and concentration so that the thermodynamic state of the mixture in the control volume can be defined.
Fig. 2. Predicted pressure difference between suction cavity and suction plenum (injection to suction plenum (black) or during the start phase of compression (grey)).
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4. Injection into suction versus injection into compression cavity In the previous experiments by Zaytsev [4] all the injection flow was introduced into the suction plenum (volume between inlet nozzle and first compressor cavity). As a conclusion of those experiments Zaytsev recommended avoiding injection into the suction plenum due to the large vapor flashing that is produced in this way. In the new prototype only one port is used for injection. This port is located on the male rotor side at a shaft angle of 432.58, which corresponds to the start phase of compression. The injection will take place for 728, the phase shift between rotor grooves. To understand the effect of injection into the suction plenum, Zaytsev’s [4] thermodynamic compressor simulation model has been used. Fig. 2 shows a plot of the predicted difference between the pressure in the suction line and the pressure in the suction cavity for the operating conditions listed in Table 1. For almost the whole suction phase the pressure in the suction cavity is slightly higher than the pressure in the suction line. Injection in the suction plenum leads to a slight increase of the pressure during the suction process. The suction plenum has a large flow passage area and works with a very small pressure drop between plenum and cavity. Any slight increase of pressure in that space reduces the suction flow rate dramatically. Practically almost no ‘new’ suction flow is present in the suction cavity and mainly the leakage flow flashing into the suction cavity is re-compressed. Fig. 2 also shows the model results when the suction liquid plus labyrinth leakage flow are injected during the start phase of compression. The pressure in the suction cavity is lower than the pressure in the suction line during the largest portion of the suction phase. Consequently, the mass flow passing through the suction port is mostly positive and the model predicts a much better performance for the compressor. Table 1 gives an overview of the performance indicators. The isentropic efficiency increases from 7 to 51%, while the volumetric efficiency increases from 7 to 57% and the power consumption reduces from 12.2 to 10.9 kW as the injection point is moved from suction plenum to the start phase of compression. An appropriate switching of the valves allows the injection flow to be directed to the suction plenum port ‘S’ or to the injection port ‘I’, see Fig. 1. In this way the impact of the position at which injection takes place on performance can be experimentally verified. Table 2 compares experimental results of injection in the suction plenum with injection during the start phase of compression for a fixed pressure ratio of 3.6. When the liquid is injected into the suction plenum, the isentropic efficiency of the compressor is very low as also predicted by Zaytsev’s model. The isentropic efficiency decreases slightly when the suction pressure increases for a fixed pressure ratio, experiment 23 (pSuctZ1.71 bar;
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Table 1 Impact of position of injection on performance indicators (model)
x0 (kg/kg) pSuct (bar) pDisch/pSuct TSuct (C) TDisch (C) m_ GSuct (kg/s) m_ LSuct (kg/s) m_ Disch (kg/s) his hv P (kW)
Injection to suction
Injection at 432.58
0.376 3.7 2.5 62.8 96.9 0.006 0.057 0.078 0.070 0.074 12.2
0.376 3.7 2.5 62.8 82.5 0.050 0.459 0.510 0.510 0.571 10.9
hisZ9%) versus 22 (pSuctZ2.34 bar; hisZ8%). When the suction pressure increases, for a fixed pressure ratio, the pressure difference across the compressor pDischKpSuct increases too. This explains the slight decrease of the compressor performance with increased suction pressure. The isentropic efficiency of the compressor is about 35% when the liquid is injected during the start phase of compression. The differences encountered between theoretical and experimental isentropic efficiency in case of injection during the start phase of compression result from the amount of liquid injected: the higher efficiencies of the model are obtained with liquid injection flows of 459 g sK1 while the experimental lower efficiencies are obtained with injection flows of 30–34 g sK1. This leads to higher compressor discharge temperatures, see Tables 1 and 2, and so to larger irreversibility losses.
used to identify the ideal location of the injection port. The results are shown in Fig. 3 where the isentropic efficiency is shown as a function of the male rotor rotation angle. The simulation program takes into account the irreversibilities due to leakage flows, discharge and suction pressure drop. The labyrinth leakage flow was re-injected into the same port as the liquid injection flow. Fig. 3 was obtained from simulations where 41 start injection was taken as 384, 456, 528, 600, 672, and 41 stop injectionZ 41 start injectionC728. When injecting at higher pressures, the labyrinth leakage mass flow rate reduces. The negative effect is that the pressure on the discharge side lip seal increases so that friction losses and wear rate increase. The ideal location of the injection port is during the start phase of compression. The location of the injection port for the experiments reported in the previous section is 432.58 so that injection takes place at the ideal position.
5. Impact of liquid injection angle
6. Labyrinth seal leakage
Compression starts at a male shaft rotation angle of 3848. The thermodynamic compressor simulation model has been
Fig. 4 presents the discharge end-face of a screw compressor that, similarly to the experimental compressor,
Table 2 Impact of position of injection on performance indicators (experiments) Estimated maximum uncertainty (%) X0 (kg/kg) pSuct (bar) pDisch/pSuct TSuct (C) TDisch (C) m_ GSuct (kg/s) m_ LSuct (kg/s) m_ Disch (kg/s) his hv P (kW)
6 8 12 1 1 1 1 1 18 12 4
Injection to suction
Injection at 432.58
Exp. 22
Exp. 23
Exp. 11
Exp. 13
0.350 2.34 3.6 65.8 133.3 0.002 0.016 0.018 0.080 0.090 11.4
0.350 1.71 3.6 50.3 110.4 0.002 0.016 0.019 0.090 0.100 7.8
0.340 1.02 3.6 38.8 88.6 0.004 0.030 0.034 0.360 0.370 4.8
0.340 1.26 3.6 42.4 89.8 0.004 0.034 0.038 0.340 0.290 5.8
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Fig. 3. Isentropic efficiency as a function of the injection port angle.
has five lobes (Zz1) on the male rotor. One important leakage path—illustrated with white arrows—is formed between any leading cavity and the corresponding trailing cavity. The pressures of leading and trailing cavities have a phase shift equal to D4Z2p/z1. Assume that in an annular region around the labyrinth seal’s entrance—as indicated with the notation UPR (uniform property region) in Fig. 4— the working fluid has uniform thermodynamic properties [13]. This assumption is reasonable if it is considered that the fluid at the entrance of the labyrinth seal’s clearance (close to the shaft perimeter) is well mixed due to the movement of the rotors. This allows for the construction of an equivalent network of leakage paths, as presented in Fig. 5. The cavities being at a more advanced compression Fig. 5. Leakage model for the discharge end-face of screw compressors.
phase deliver fluid to the UPR. By contrary, the cavities being at the beginning of the compression process, receive fluid from the UPR while they have a lower pressure. The flow also leaks from the UPR through the labyrinth seal clearance and it is eventually re-injected into the process at a lower pressure, pinj. The leakage flows from leading to trailing cavities are not represented in the network of Fig. 5. These leakages are treated separately and are superimposed on the ones derived with the UPR model. The simplest flow model widely used in simulation of screw compressors is the converging nozzle flow model, with assumptions that the compressible flow is isentropic and the pressure in the narrowest part of the flow path is equal to the downstream cavity pressure—see Zaytsev [4], Taniguchi et al. [14]. Although the non-isothermal phase change complicates the mathematical treatment, the converging nozzle flow model can also be used for nonazeotropic refrigerants (such as ammonia–water). The calculated flow rate is adjusted by an empirical flow coefficient, C0: Fig. 4. Flow leakage paths at the discharge end-face, and the UPR around the labyrinth.
_ 1 Þ Z C0 rAef W 41 mð4
(4)
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The flow velocity, W, is bounded by the local speed of sound and takes viscous effects into account. The flow coefficient is usually found from experiments. With reference to Fig. 4 the end-face leakage area between any cavity and the UPR is assumed to be given by:
Aef ð41 Þ Z lAB C lCD 41 def (5)
gives: ( ALab
Z
pUPR vUPR N
z1 X
rAef W
iZ1
Where lAB and lCD have maximum values when both tips of the lobes, that form the cavity, are directed to the housing. This situation stands for almost all compression and discharge phases, except for the short period when the male and female lobes are in contact with each other. In that period the leakage area between the cavity and the UPR decreases to zero; this reduction is assumed to by linear. The position of arcs AB and CD is considered at an average radius between those corresponding to the labyrinths and rotors foot circles. The mass conservation in the UPR, based on the leakage network of Fig. 5, is m_ Lab ð41 Þ Z
z1 X
m_
iZ1
41CðiK1Þ 2p z
(6)
1=2 1=2 ) pinj 2 1K pUPR 4
(9)
41 CðiK1Þ 2p z 1
Assuming that pinj and the thermodynamic state at all compression cavities (iZ1.z1) are known, the only unknown in Eq. (9) remains the thermodynamic state of the UPR. This is defined by three parameters: the temperature, TUPR, the pressure, pUPR, and the overall concentration xUPR. A second equation is the mass conservation of ammonia ( 1=2 ) pinj 2 pUPR 1=2 xUPR ALab 1K vUPR N pUPR 4 1
Z
z1 X
rAef Wx0
iZ1
1
1
(10)
41 CðiK1Þ 2p z 1
The left hand side of Eq. (6) refers to the mass flow rate that leaks through the labyrinth seal. The prediction of leakage mass flow rate through labyrinth seals has been studied extensively in turbo machinery and compressor engineering. A comparison between the leakage flow rate predicted with the models proposed by Eser and Kazakia [15], Mu¨ller and Nau [16], Yucel and Kazakia [17], Stodola et al. [18] and Zaytsev [4] indicates that the model by Stodola et al. combines simplicity with reasonable accuracy. This model has been adopted: 1=2 1=2 pinj 2 p m_ Lab ð41 Þ Z C0 ALab UPR 1K (7) vUPR N pUPR
The third equation is the energy conservation law of all z1 leakage flows that enter or leave the UPR at the end-face of the rotors of the screw compressor. The flow velocity through the end-face clearance is high so that heat transfer between rotors and housing can be neglected. The conservation of energy of each of the z1 flows can be written as:
Where pinj represents the pressure at the injection port, where the fluid that leaks through the labyrinth seal is reinjected into the compression process. The mass flow depends on the annular passage area of the labyrinth seal, ALab, and the number of labyrinth blades, N. The passage area is approximated with
i Z 1.z1
ALab Z pðD1 d1 C D2 d2 Þ
(8)
Furthermore, it is assumed that the flow coefficient C0 has the same value for both labyrinth flow and end-face leakage flows. In both cases, C0 takes the entropy production during the expansion process into account. Substitution
hUPR ð41 Þ C 0:5W 2
UPR 41 CðiK1Þ 2p z
1
Z ðh C 0:5W 2 Þ
41 CðiK1Þ 2p z
;
(11)
1
where WUPR is the flow velocity at the UPR boundary, while W is the flow velocity at the working cavity boundary. Expressing the velocities as function of the mass flow rate, flow passage area, and the specific volume: " # 1=2 pinj 2 pUPR 1=2 ALab 1K hUPR vUPR N pUPR 4 1
z1 X 2 Z rAef W h C 0:5 W 2 KWUPR 41CðiK1Þ 2p z1
iZ1
(12)
Table 3 Comparison between measured and computed leakage flow through the labyrinth seals Method
Measured/computed leakage flow (g/s)
Percentage difference with measured flow rate (%)
Experiment Mu¨ller and Nau [16] Stodola et al. [18]
7.2 7.1 7.6
– 1.2 6.2
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The fluid enthalpy in the UPR, hUPR, is related through the equation of state to the fluid temperature, TUPR and pressure, pUPR. For a two-component homogeneous mixture with overall concentration xUPR: hUPR Z h pUPR ; TUPR ; xUPR (13) A Newton–Raphson method can be used to solve the system of equations for pUPR, TUPR, xUPR. The experimental set-up made it possible to measure the labyrinth leakage flow. A comparison of the measured flow rate with the computed flow rate, for one operating condition, is given in Table 3. The inlet temperature was 73 8C and the concentration of the flow through the labyrinth seal was 0.357, which is the overall concentration in the system. The pressure at the outlet of the seal is equal to 3.72 bar. The average pressure at the inlet of the seal, calculated as the average between discharge and injection pressure, is 5.05 bar. This assumption is reasonable because all five cavities of the rotor are connected to the discharge end-face. The pressure during the start of compression is equal to the injection pressure (the liquid is injected during the start phase of compression) and the pressure in the last cavity is equal to the discharge pressure. The suction port is also connected to the discharge end-face but only for a short period (shaft angles between 314 and 3848). Table 3 shows that Mu¨ller and Nau’s [16] method gives the best prediction of the leakage flow through the labyrinth seal. The prediction by Stodola et al.’s method [18] differs by only 6% from the measured leakage flow. It may be concluded that both Mu¨ller and Nau’s and Stodola et al.’s methods reasonably predict the leakage flow through the labyrinth seal. The added advantage of Stodola et al.’s Eq. (7) is its simplicity. The injection flow rate is for this experiment ca. 23 g sK1. When the measured leakage flow is compared with the amount of liquid that is injected, it can be seen that about 31% of the mass leaks out through the labyrinth seals. This indicates that the leakage flow through the labyrinth seals is very large. It can be expected that it will lead to a substantial decrease of the isentropic efficiency because a significant
Fig. 6. Compressor indicated diagrams: predicted versus measured.
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amount of work was spent to compress the leaking working fluid. Fig. 6 shows an experimental indicated diagram together with the result of simulations performed using Eq. (7) derived by Stodola et al. [18] for computing the leakage flow through the labyrinth seal. To fit the predicted p–V indicated diagrams to the experimentally determined indicated diagrams a discharge flow coefficient CD and a leakage paths flow coefficient C0 have been used. The value used for the discharge flow coefficient CD was 0.8, in agreement with the values reported by Taniguchi et al. [14]. If the pressure at the outlet of a converging nozzle is considered to correspond with the pressure at the nozzle then the flow coefficient C0 should have a value lower than one. The best fit is obtained for a flow coefficient of 1.2. The value of 1.2 for the flow coefficient implies that the outlet pressure condition should not be applied to the narrowest gap but to a location downstream of this location. Alternatively, the large flow coefficient may result from the size of the leakage paths as the compressor runs and/or the density of the leakage flow. The agreement between model and experimental indicated diagram is very good. Simulations with the thermodynamic model indicate that, departing from an operating condition with isentropic efficiency of 6% and maintaining all other irreversibilities unmodified, the isentropic efficiency increases to 18% when the labyrinth leakage flow is reduced with 50%. Methods to reduce this leakage flow can be derived from Eqs. (7) and (8).
7. Impact of injection flow The compressor thermodynamic simulation model [4] has been used to investigate the influence of the vapor quality of the leakage flow through the rotor tips on the compressor performance. Fig. 7 shows the simulation results for vapor quality of the leakage flow, xLK, in the range 0– 0.3, representing liquid only to mainly vapor conditions. Results for homogeneous conditions give an isentropic efficiency of 5.4% and are indicated with an arrow. This value corresponds to an experimentally obtained isentropic
Fig. 7. The sensitivity of isentropic efficiency to vapor quality of leakage flow.
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efficiency. These results demonstrate qualitatively and quantitatively the sensitivity of the isentropic efficiency to the vapor quality of the leakage flow through the rotors tip clearances. Additional liquid in the leakage paths (even in small quantities) affects the result of the simulation. By contrary, a larger presence of vapor in leakage paths has a negligible effect on the simulation results. When liquid flashes to lower pressures, a large portion of the leakage flow still remains in the liquid phase. The liquid is much easier to transport to higher pressures than the vapor, which must be re-compressed by spending additional work. Qualitatively, larger liquid amounts in leakage paths lead to higher compressor isentropic efficiencies. During the experiments the injected liquid flow has been varied. In most experiments the liquid was injected from the suction side but for a few experiments extra liquid has been injected from the discharge side. Fig. 8 shows the indicated isentropic efficiency of all the experiments with liquid injection during the start phase of compression. Also the experiments at a different rotational speed and pressure ratio are included leading to some scatter. As expected, there is a minimum quantity of injected liquid required to obtain reasonable efficiencies. Since the suction liquid injection pump had a limited capacity, the maximum flow that could be attained from suction was about 35 g sK1. As Fig. 8 indicates, higher efficiencies require larger liquid injection flows. To further identify the effect of larger liquid injection flows, liquid from the highpressure separator was added to the suction side flow. Before being added, the liquid was subcooled to prevent flashing, see Fig. 1 for injection scheme. These experiments are indicated with triangles and, as indicated by the trend line, re-injection from the high pressure side leads to a slight decrease in efficiency but to a similar trend for the isentropic efficiency. The trend shown in Fig. 8 indicates that the isentropic efficiency of the compressor could still significantly be improved by increasing the liquid injection flow. This flow should preferably come from the suction side of the
Fig. 8. Indicated isentropic efficiency as a function of the total injected liquid flow.
compressor but also liquid injected from the high pressure side has a very positive effect. 8. Conclusions Both model and experiments show a significant impact of the liquid injection location. When the injection location is moved from the suction plenum to the start phase of compression, the model gives indicated isentropic efficiency increases from 5 to 50% while the experiments show an increase from 10 to 35%. The lower increase for the experiments most probably results from the smaller amount of injected liquid: 30 to 35 g sK1 in the experiments against 459 g sK1 in the model. The ideal location for liquid injection is during the start phase of compression. The labyrinth leakage flow is substantial and has a very large impact on the compressor performance. Both model and experiments show a significant effect of liquid injection on compressor performance. Also liquid injection from the high pressure side leads to significant improvement of the compressor performance. Acknowledgements This study was conducted as part of the NOVEM BSENEO Programme (BSE-NEO 0268.02.03.03.0002) and was partly financed by NOVEM.
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C.A.I. Ferreira et al. / International Journal of Refrigeration 29 (2006) 556–565 [8] W. Malewski, Integrated absorption and compression heat pump cycle using mixed working fluid ammonia and water, Proceedings of the 2nd International Workshop on Research Activities on Advanced Heat Pumps, University of Technology Graz, Graz, Austria, 1988. p. 35–44. [9] G. Bergmann, G. Hivessy, Main features and operational experience of the hybrid heat pump pilot plant, Proceedings of the 3rd International Workshop on Research Activities on Advanced Heat Pumps, University of Technology Graz, Graz, Austria, 1990. p. 111–120. [10] C. Zamfirescu, N. Nannan, M. Marin, C.A. Infante Ferreira, Oil free two phase ammonia (water) compressor. Final report of NOVEM project BSE-NEO 0268.02.03.03.0002, Delft University of Technology, Delft; 2004 [report K-336]. [11] B. Ziegler, Ch. Trepp, Equation of state for ammonia–water mixtures, Int J Refrigeration 7 (1984) 101–106. [12] D. Zaytsev, C.A. Infante Ferreira, Profile generation method for twin screw compressor rotors based on the meshing line, Int J Refrigeration 28 (2005) 744–755.
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[13] C. Zamfirescu, C.A. Infante Ferreira, A uniform property region method for screw compressor’s end face leakage prediction, Proceedings of the 17th International Compressor Engineering Conference at Purdue; 2004, Purdue University, West Lafayette, 2004. [paper C-142]. [14] H. Taniguchi, K. Kudo, W.H. Giedt, I. Park, S. Kumazawa, Analytical and experimental investigation of the two-phase flow screw expanders for power generation, J Eng Gas Turb Power 110 (1988) 628–635. [15] D. Eser, J.Y. Kazakia, Air flow in cavities of labyrinth seals, Int J Eng Sci 33 (15) (1995) 2309–2326. [16] H.K. Mu¨ller, B.S. Nau, Fluid sealing technology, Marcel Dekker Inc., New York, 1998. [17] U. Yucel, J. Kazakia, Analytical prediction techniques for axisymmetric flows in gas labyrinth seals’, J Eng Gas Turb Power 123 (2001) 255–257. [18] Stodola, Gercke, Traupel. Cited in: Trutnovsky K, Komotori K. Beru¨hrungsfreie Dichtungen. VDI-Verlag GmbH, Du¨sseldorf; 1981 [in German].
Twin screw oil-free wet compressor for compression–absorption cycle C.A. Infante Ferreira1,*, C. Zamfirescu, D. Zaytsev2 Mechanical, Maritime and Materials Engineering, Delft University of Technology, Mekelweg 2, 2628 CD Delft, The Netherlands Received 10 February 2005; received in revised form 15 September 2005; accepted 10 October 2005 Available online 20 December 2005
Abstract The performance of a twin screw compressor operating under wet (two-phase) compression conditions in an ammonia–water compression absorption heat pump cycle is investigated both theoretically and experimentally. The paper reports on the influence of the location of liquid intake or, depending what applies, injection angle and mass flow rate of the injected liquid on compressor performance. Labyrinth seals separate the oil-free process side from oil lubricated bearing housing. Labyrinth seals leakage is modelled and its impact on performance is theoretically and experimentally investigated. The need for liquid injection from the discharge side to obtain acceptable performance is discussed based on experimental results. q 2005 Elsevier Ltd and IIR. All rights reserved. Keywords: Absorption system; Ammonia-water; Screw compressor; Oil-free compressor; Research; Experiment; Performance
Compresseur a` double vis secs dans un cycle a` compression a` absorption Mots cle´s : Syste`me a` absorption ; Ammoniac-eau ; Compresseur a` vis ; Compresseur sec ; Recherche ; Expe´rimentaion ; Performance
1. Introduction Compression–absorption heat pump cycles have the potential to significantly improve the performance of heat pump cycles based on the vapor compression principle. Itard [1,2], Hulte´n and Berntsson [3] and more recently Zaytsev
* Corresponding author. Tel.: C31 152784894; fax: C31 152782460. E-mail address: [email protected] (C.A.I. Ferreira). 1 Member of IIR Commission B1. 2 Present address: Grasso GmbH RT, R&D Screw Compressors, Holzhauser Strasse 165, 13509 Berlin, Germany.
0140-7007/$35.00 q 2005 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2005.10.006
[4] investigated this potential. For high temperature industrial heat pumps, a gain of about 20% in comparison with ‘conventional’ vapor compressor heat pumps can be attained. Two different ways of implementing compressor– absorption heat pump cycles have been investigated in the past: the wet compression and the dry compression types. The differences between the two solutions have extensively been discussed by Itard [1,2]. The dry compression implementation raises two problems: (i) the irreversibility losses associated with vapor superheating play a major role in degrading the overall performance of the cycle; (ii) dry compressors require oil lubrication, with the implication of oil contamination of the liquid refrigerant [5] and additional degradation of the performance. Since the liquid refrigerant
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Nomenclature Aef ALab CD C0 D h j l m m_ n N p Q T u v
leakage flow area (m2) labyrinth leakage flow area (m2) empirical discharge flow coefficient empirical leakage paths flow coefficient labyrinth blade diameter (m) enthalpy (J kgK1) number of inflow paths length (m) mass (kg) mass flow (kg sK1) number of outflow paths number of labyrinth blades pressure (Pa) heat (J) temperature (K) internal energy (J kgK1) specific volume (m3 kgK1)
and the oil have similar densities, very expensive separation devices must be included in the heat pump cycle. Oil-free compressors operating under wet conditions avoid superheating. The liquid refrigerant is used as lubricant and the oil—still needed for lubrication of the bearings—is separated from the compression process. The performance advantage of the wet compression in comparison with the dry compression varies from 2.5 to 13.0% for the application cases studied by Itard [1]. However, the lack of information related to two-phase compression processes, and the use of liquid refrigerant itself as a lubricant, makes the development of a suitable ‘wet compressor’ difficult. Zaytsev [4] gives an overview of compression–absorption cycle investigations that have been carried out in recent years. Most investigations have been dedicated to the dry compression type. Very recently, Satapathy et al. [6] and Risberg et al. [7] reported on dry compression systems. Malewski [8], Bergmann and Hivessy [9], Itard [1] and Zaytsev [4] have reported experimental studies on wet compression absorption cycles. Zaytsev [4] reviewed the different compressor types available in the market for their applicability as compressor for wet compression absorption cycles. Considering the specific requirements of these cycles, the screw compressors appeared to be the most suitable choice. They are tolerant to liquid carry over and have rather high efficiency. Zaytsev modified an existing twin-screw compressor available on the market in order to meet the requirements imposed by the wet compression absorption application, implemented it in an experimental compression absorption heat pump and collected several sets of experimental data with it. The isentropic efficiency of this first prototype was rather
V W x z
volume of cavity (m3) velocity (m sK1) ammonia concentration number of rotor’s lobes
Greek letters d clearance (m) r density (kg mK3) 4 rotation angle Subscripts 0 overall 1 male rotor 2 female rotor ef effective inj injection port Lab labyrinth UPR uniform property region
low because the liquid refrigerant was injected into the suction cavity where it flashed into vapor, reducing the volume of displaced vapor significantly. Zaytsev [4] suggested that the compressor efficiency might be improved if the low pressure liquid is separated from the vapor and injected, together with the labyrinth leakage flow (needed to allow for oil-free operation), into the cavity at the start phase of compression. This paper reports the experimental and theoretical investigations carried out to more accurately understand and model the physical processes related to wet compression. Aim is to increase the isentropic efficiency of the compressor to a level that makes wet compression– absorption heat pumps a competitive technical option. Further details of this investigation can be found in Zamfirescu et al. [10]. 2. Experimental set-up The experimental data collected by Zaytsev [4] allowed to re-design the rotors taking thermal expansion into account. The newly built compressor prototype has reduced clearances (at rotor tips, discharge end face, and labyrinth seal), and only one injection port. The experimental ammonia–water compression–absorption set-up previously used by Zaytsev [4] has been used after slight modifications, see Fig. 1. It allows for experimental determination of the compressor performance indicators including indicated diagrams. Also data related to the labyrinth leakage flow have been measured. For data processing purpose the heat pump was split in three functional parts delimited by three states defined as follows: (i) state A is a combination of states A1 and A2, i.e. compressor suction and liquid injection from low pressure
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Fig. 1. Experimental set-up with sample experimental data.
side (at flow meter ‘FT4’); (ii) state B represents the absorber inlet; the properties in state B are a combination of the properties in the liquid line (at mass flow meter ‘mFT5’) and in the vapor line; (iii) state C is located at the absorber outlet. Six Kistler high-frequency piezo-electric pressure transducers in combination with an angle transducer have been used for measurement of the compressor pV-diagram. The first pressure transducer is situated 98 before the compression start point, while the last sensor is placed 258 after the discharge start point. The compressor shaft power is obtained from the driving torque (Koyowa torque transducer) and shaft speed measurements. Further details of the experimental set-up, estimated accuracy of the measurements and resulting uncertainties for relevant parameters can be found in [4]. Table 2 includes a column where the maximum experimental uncertainties of the considered performance indicators are given. The large uncertainty for the efficiencies originates mainly from the calculation of the ammonia concentration.
3. Thermodynamic model Two options of modeling the wet compression of ammonia–water solution were analyzed: a pu-model and a pT-model. Due to the high nonlinearity of the equation of state, the pu-model requires a time consuming iterative procedure. Although the governing equations of the pTmodel are more complicated than those of the pu-model, with the existing ammonia–water subroutines based on Ziegler–Trepp [11] the computation time needed to integrate the pT-model governing equations is shorter. The pressure and temperature are the input variables of the mixture thermodynamic state model; therefore, there is no need of additional iterations to calculate the rest of the mixture state parameters. The model assumes that liquid and vapor are at thermodynamic equilibrium at any moment and that the thermodynamic state of the solution is given by a single set of governing equations (homogeneous model). The
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governing equations for the control volume (cavity) under study are (see [4,12] for further details): The mass conservation equation: dp 1 Z vv d41
vp T;x
C
"
n X
v m
kZ1
dmout d41
! j X dmin K d41 k k kZ1
1 dV vv dT vv dx0 : K K m d41 vT p;x d41 vx0 p;T d41
(1)
The ammonia mass conservation equation: " # j j X dx0 1 X dmin dmin Z x Kx0 : m kZ1 0in;k d41 k d41 d41 k kZ1
(2)
The energy conservation equation: dT Z d41 T
vv
v vT p;x m
n P dmout kZ1
d41
k
K
kZ1 vv vh vp T;x vT
dQ d41
C
C
j P dmin d41
k
p;x C T
C m1
vv 2
vT p;x
vh in ðhin;k KhÞ dm d41 k Km vx0 p;T kZ1 vh vv 2 mT vT m vT p;x C vv p;x j P
dx0 dV vv d41 K vx0 p;T d41
dx0 d41
:
ð3Þ
vp T;x
A numerical integration of the governing equations is used to calculate the three thermodynamic parameters of the mixture: pressure, temperature and concentration so that the thermodynamic state of the mixture in the control volume can be defined.
Fig. 2. Predicted pressure difference between suction cavity and suction plenum (injection to suction plenum (black) or during the start phase of compression (grey)).
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4. Injection into suction versus injection into compression cavity In the previous experiments by Zaytsev [4] all the injection flow was introduced into the suction plenum (volume between inlet nozzle and first compressor cavity). As a conclusion of those experiments Zaytsev recommended avoiding injection into the suction plenum due to the large vapor flashing that is produced in this way. In the new prototype only one port is used for injection. This port is located on the male rotor side at a shaft angle of 432.58, which corresponds to the start phase of compression. The injection will take place for 728, the phase shift between rotor grooves. To understand the effect of injection into the suction plenum, Zaytsev’s [4] thermodynamic compressor simulation model has been used. Fig. 2 shows a plot of the predicted difference between the pressure in the suction line and the pressure in the suction cavity for the operating conditions listed in Table 1. For almost the whole suction phase the pressure in the suction cavity is slightly higher than the pressure in the suction line. Injection in the suction plenum leads to a slight increase of the pressure during the suction process. The suction plenum has a large flow passage area and works with a very small pressure drop between plenum and cavity. Any slight increase of pressure in that space reduces the suction flow rate dramatically. Practically almost no ‘new’ suction flow is present in the suction cavity and mainly the leakage flow flashing into the suction cavity is re-compressed. Fig. 2 also shows the model results when the suction liquid plus labyrinth leakage flow are injected during the start phase of compression. The pressure in the suction cavity is lower than the pressure in the suction line during the largest portion of the suction phase. Consequently, the mass flow passing through the suction port is mostly positive and the model predicts a much better performance for the compressor. Table 1 gives an overview of the performance indicators. The isentropic efficiency increases from 7 to 51%, while the volumetric efficiency increases from 7 to 57% and the power consumption reduces from 12.2 to 10.9 kW as the injection point is moved from suction plenum to the start phase of compression. An appropriate switching of the valves allows the injection flow to be directed to the suction plenum port ‘S’ or to the injection port ‘I’, see Fig. 1. In this way the impact of the position at which injection takes place on performance can be experimentally verified. Table 2 compares experimental results of injection in the suction plenum with injection during the start phase of compression for a fixed pressure ratio of 3.6. When the liquid is injected into the suction plenum, the isentropic efficiency of the compressor is very low as also predicted by Zaytsev’s model. The isentropic efficiency decreases slightly when the suction pressure increases for a fixed pressure ratio, experiment 23 (pSuctZ1.71 bar;
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Table 1 Impact of position of injection on performance indicators (model)
x0 (kg/kg) pSuct (bar) pDisch/pSuct TSuct (C) TDisch (C) m_ GSuct (kg/s) m_ LSuct (kg/s) m_ Disch (kg/s) his hv P (kW)
Injection to suction
Injection at 432.58
0.376 3.7 2.5 62.8 96.9 0.006 0.057 0.078 0.070 0.074 12.2
0.376 3.7 2.5 62.8 82.5 0.050 0.459 0.510 0.510 0.571 10.9
hisZ9%) versus 22 (pSuctZ2.34 bar; hisZ8%). When the suction pressure increases, for a fixed pressure ratio, the pressure difference across the compressor pDischKpSuct increases too. This explains the slight decrease of the compressor performance with increased suction pressure. The isentropic efficiency of the compressor is about 35% when the liquid is injected during the start phase of compression. The differences encountered between theoretical and experimental isentropic efficiency in case of injection during the start phase of compression result from the amount of liquid injected: the higher efficiencies of the model are obtained with liquid injection flows of 459 g sK1 while the experimental lower efficiencies are obtained with injection flows of 30–34 g sK1. This leads to higher compressor discharge temperatures, see Tables 1 and 2, and so to larger irreversibility losses.
used to identify the ideal location of the injection port. The results are shown in Fig. 3 where the isentropic efficiency is shown as a function of the male rotor rotation angle. The simulation program takes into account the irreversibilities due to leakage flows, discharge and suction pressure drop. The labyrinth leakage flow was re-injected into the same port as the liquid injection flow. Fig. 3 was obtained from simulations where 41 start injection was taken as 384, 456, 528, 600, 672, and 41 stop injectionZ 41 start injectionC728. When injecting at higher pressures, the labyrinth leakage mass flow rate reduces. The negative effect is that the pressure on the discharge side lip seal increases so that friction losses and wear rate increase. The ideal location of the injection port is during the start phase of compression. The location of the injection port for the experiments reported in the previous section is 432.58 so that injection takes place at the ideal position.
5. Impact of liquid injection angle
6. Labyrinth seal leakage
Compression starts at a male shaft rotation angle of 3848. The thermodynamic compressor simulation model has been
Fig. 4 presents the discharge end-face of a screw compressor that, similarly to the experimental compressor,
Table 2 Impact of position of injection on performance indicators (experiments) Estimated maximum uncertainty (%) X0 (kg/kg) pSuct (bar) pDisch/pSuct TSuct (C) TDisch (C) m_ GSuct (kg/s) m_ LSuct (kg/s) m_ Disch (kg/s) his hv P (kW)
6 8 12 1 1 1 1 1 18 12 4
Injection to suction
Injection at 432.58
Exp. 22
Exp. 23
Exp. 11
Exp. 13
0.350 2.34 3.6 65.8 133.3 0.002 0.016 0.018 0.080 0.090 11.4
0.350 1.71 3.6 50.3 110.4 0.002 0.016 0.019 0.090 0.100 7.8
0.340 1.02 3.6 38.8 88.6 0.004 0.030 0.034 0.360 0.370 4.8
0.340 1.26 3.6 42.4 89.8 0.004 0.034 0.038 0.340 0.290 5.8
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Fig. 3. Isentropic efficiency as a function of the injection port angle.
has five lobes (Zz1) on the male rotor. One important leakage path—illustrated with white arrows—is formed between any leading cavity and the corresponding trailing cavity. The pressures of leading and trailing cavities have a phase shift equal to D4Z2p/z1. Assume that in an annular region around the labyrinth seal’s entrance—as indicated with the notation UPR (uniform property region) in Fig. 4— the working fluid has uniform thermodynamic properties [13]. This assumption is reasonable if it is considered that the fluid at the entrance of the labyrinth seal’s clearance (close to the shaft perimeter) is well mixed due to the movement of the rotors. This allows for the construction of an equivalent network of leakage paths, as presented in Fig. 5. The cavities being at a more advanced compression Fig. 5. Leakage model for the discharge end-face of screw compressors.
phase deliver fluid to the UPR. By contrary, the cavities being at the beginning of the compression process, receive fluid from the UPR while they have a lower pressure. The flow also leaks from the UPR through the labyrinth seal clearance and it is eventually re-injected into the process at a lower pressure, pinj. The leakage flows from leading to trailing cavities are not represented in the network of Fig. 5. These leakages are treated separately and are superimposed on the ones derived with the UPR model. The simplest flow model widely used in simulation of screw compressors is the converging nozzle flow model, with assumptions that the compressible flow is isentropic and the pressure in the narrowest part of the flow path is equal to the downstream cavity pressure—see Zaytsev [4], Taniguchi et al. [14]. Although the non-isothermal phase change complicates the mathematical treatment, the converging nozzle flow model can also be used for nonazeotropic refrigerants (such as ammonia–water). The calculated flow rate is adjusted by an empirical flow coefficient, C0: Fig. 4. Flow leakage paths at the discharge end-face, and the UPR around the labyrinth.
_ 1 Þ Z C0 rAef W 41 mð4
(4)
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The flow velocity, W, is bounded by the local speed of sound and takes viscous effects into account. The flow coefficient is usually found from experiments. With reference to Fig. 4 the end-face leakage area between any cavity and the UPR is assumed to be given by:
Aef ð41 Þ Z lAB C lCD 41 def (5)
gives: ( ALab
Z
pUPR vUPR N
z1 X
rAef W
iZ1
Where lAB and lCD have maximum values when both tips of the lobes, that form the cavity, are directed to the housing. This situation stands for almost all compression and discharge phases, except for the short period when the male and female lobes are in contact with each other. In that period the leakage area between the cavity and the UPR decreases to zero; this reduction is assumed to by linear. The position of arcs AB and CD is considered at an average radius between those corresponding to the labyrinths and rotors foot circles. The mass conservation in the UPR, based on the leakage network of Fig. 5, is m_ Lab ð41 Þ Z
z1 X
m_
iZ1
41CðiK1Þ 2p z
(6)
1=2 1=2 ) pinj 2 1K pUPR 4
(9)
41 CðiK1Þ 2p z 1
Assuming that pinj and the thermodynamic state at all compression cavities (iZ1.z1) are known, the only unknown in Eq. (9) remains the thermodynamic state of the UPR. This is defined by three parameters: the temperature, TUPR, the pressure, pUPR, and the overall concentration xUPR. A second equation is the mass conservation of ammonia ( 1=2 ) pinj 2 pUPR 1=2 xUPR ALab 1K vUPR N pUPR 4 1
Z
z1 X
rAef Wx0
iZ1
1
1
(10)
41 CðiK1Þ 2p z 1
The left hand side of Eq. (6) refers to the mass flow rate that leaks through the labyrinth seal. The prediction of leakage mass flow rate through labyrinth seals has been studied extensively in turbo machinery and compressor engineering. A comparison between the leakage flow rate predicted with the models proposed by Eser and Kazakia [15], Mu¨ller and Nau [16], Yucel and Kazakia [17], Stodola et al. [18] and Zaytsev [4] indicates that the model by Stodola et al. combines simplicity with reasonable accuracy. This model has been adopted: 1=2 1=2 pinj 2 p m_ Lab ð41 Þ Z C0 ALab UPR 1K (7) vUPR N pUPR
The third equation is the energy conservation law of all z1 leakage flows that enter or leave the UPR at the end-face of the rotors of the screw compressor. The flow velocity through the end-face clearance is high so that heat transfer between rotors and housing can be neglected. The conservation of energy of each of the z1 flows can be written as:
Where pinj represents the pressure at the injection port, where the fluid that leaks through the labyrinth seal is reinjected into the compression process. The mass flow depends on the annular passage area of the labyrinth seal, ALab, and the number of labyrinth blades, N. The passage area is approximated with
i Z 1.z1
ALab Z pðD1 d1 C D2 d2 Þ
(8)
Furthermore, it is assumed that the flow coefficient C0 has the same value for both labyrinth flow and end-face leakage flows. In both cases, C0 takes the entropy production during the expansion process into account. Substitution
hUPR ð41 Þ C 0:5W 2
UPR 41 CðiK1Þ 2p z
1
Z ðh C 0:5W 2 Þ
41 CðiK1Þ 2p z
;
(11)
1
where WUPR is the flow velocity at the UPR boundary, while W is the flow velocity at the working cavity boundary. Expressing the velocities as function of the mass flow rate, flow passage area, and the specific volume: " # 1=2 pinj 2 pUPR 1=2 ALab 1K hUPR vUPR N pUPR 4 1
z1 X 2 Z rAef W h C 0:5 W 2 KWUPR 41CðiK1Þ 2p z1
iZ1
(12)
Table 3 Comparison between measured and computed leakage flow through the labyrinth seals Method
Measured/computed leakage flow (g/s)
Percentage difference with measured flow rate (%)
Experiment Mu¨ller and Nau [16] Stodola et al. [18]
7.2 7.1 7.6
– 1.2 6.2
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The fluid enthalpy in the UPR, hUPR, is related through the equation of state to the fluid temperature, TUPR and pressure, pUPR. For a two-component homogeneous mixture with overall concentration xUPR: hUPR Z h pUPR ; TUPR ; xUPR (13) A Newton–Raphson method can be used to solve the system of equations for pUPR, TUPR, xUPR. The experimental set-up made it possible to measure the labyrinth leakage flow. A comparison of the measured flow rate with the computed flow rate, for one operating condition, is given in Table 3. The inlet temperature was 73 8C and the concentration of the flow through the labyrinth seal was 0.357, which is the overall concentration in the system. The pressure at the outlet of the seal is equal to 3.72 bar. The average pressure at the inlet of the seal, calculated as the average between discharge and injection pressure, is 5.05 bar. This assumption is reasonable because all five cavities of the rotor are connected to the discharge end-face. The pressure during the start of compression is equal to the injection pressure (the liquid is injected during the start phase of compression) and the pressure in the last cavity is equal to the discharge pressure. The suction port is also connected to the discharge end-face but only for a short period (shaft angles between 314 and 3848). Table 3 shows that Mu¨ller and Nau’s [16] method gives the best prediction of the leakage flow through the labyrinth seal. The prediction by Stodola et al.’s method [18] differs by only 6% from the measured leakage flow. It may be concluded that both Mu¨ller and Nau’s and Stodola et al.’s methods reasonably predict the leakage flow through the labyrinth seal. The added advantage of Stodola et al.’s Eq. (7) is its simplicity. The injection flow rate is for this experiment ca. 23 g sK1. When the measured leakage flow is compared with the amount of liquid that is injected, it can be seen that about 31% of the mass leaks out through the labyrinth seals. This indicates that the leakage flow through the labyrinth seals is very large. It can be expected that it will lead to a substantial decrease of the isentropic efficiency because a significant
Fig. 6. Compressor indicated diagrams: predicted versus measured.
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amount of work was spent to compress the leaking working fluid. Fig. 6 shows an experimental indicated diagram together with the result of simulations performed using Eq. (7) derived by Stodola et al. [18] for computing the leakage flow through the labyrinth seal. To fit the predicted p–V indicated diagrams to the experimentally determined indicated diagrams a discharge flow coefficient CD and a leakage paths flow coefficient C0 have been used. The value used for the discharge flow coefficient CD was 0.8, in agreement with the values reported by Taniguchi et al. [14]. If the pressure at the outlet of a converging nozzle is considered to correspond with the pressure at the nozzle then the flow coefficient C0 should have a value lower than one. The best fit is obtained for a flow coefficient of 1.2. The value of 1.2 for the flow coefficient implies that the outlet pressure condition should not be applied to the narrowest gap but to a location downstream of this location. Alternatively, the large flow coefficient may result from the size of the leakage paths as the compressor runs and/or the density of the leakage flow. The agreement between model and experimental indicated diagram is very good. Simulations with the thermodynamic model indicate that, departing from an operating condition with isentropic efficiency of 6% and maintaining all other irreversibilities unmodified, the isentropic efficiency increases to 18% when the labyrinth leakage flow is reduced with 50%. Methods to reduce this leakage flow can be derived from Eqs. (7) and (8).
7. Impact of injection flow The compressor thermodynamic simulation model [4] has been used to investigate the influence of the vapor quality of the leakage flow through the rotor tips on the compressor performance. Fig. 7 shows the simulation results for vapor quality of the leakage flow, xLK, in the range 0– 0.3, representing liquid only to mainly vapor conditions. Results for homogeneous conditions give an isentropic efficiency of 5.4% and are indicated with an arrow. This value corresponds to an experimentally obtained isentropic
Fig. 7. The sensitivity of isentropic efficiency to vapor quality of leakage flow.
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efficiency. These results demonstrate qualitatively and quantitatively the sensitivity of the isentropic efficiency to the vapor quality of the leakage flow through the rotors tip clearances. Additional liquid in the leakage paths (even in small quantities) affects the result of the simulation. By contrary, a larger presence of vapor in leakage paths has a negligible effect on the simulation results. When liquid flashes to lower pressures, a large portion of the leakage flow still remains in the liquid phase. The liquid is much easier to transport to higher pressures than the vapor, which must be re-compressed by spending additional work. Qualitatively, larger liquid amounts in leakage paths lead to higher compressor isentropic efficiencies. During the experiments the injected liquid flow has been varied. In most experiments the liquid was injected from the suction side but for a few experiments extra liquid has been injected from the discharge side. Fig. 8 shows the indicated isentropic efficiency of all the experiments with liquid injection during the start phase of compression. Also the experiments at a different rotational speed and pressure ratio are included leading to some scatter. As expected, there is a minimum quantity of injected liquid required to obtain reasonable efficiencies. Since the suction liquid injection pump had a limited capacity, the maximum flow that could be attained from suction was about 35 g sK1. As Fig. 8 indicates, higher efficiencies require larger liquid injection flows. To further identify the effect of larger liquid injection flows, liquid from the highpressure separator was added to the suction side flow. Before being added, the liquid was subcooled to prevent flashing, see Fig. 1 for injection scheme. These experiments are indicated with triangles and, as indicated by the trend line, re-injection from the high pressure side leads to a slight decrease in efficiency but to a similar trend for the isentropic efficiency. The trend shown in Fig. 8 indicates that the isentropic efficiency of the compressor could still significantly be improved by increasing the liquid injection flow. This flow should preferably come from the suction side of the
Fig. 8. Indicated isentropic efficiency as a function of the total injected liquid flow.
compressor but also liquid injected from the high pressure side has a very positive effect. 8. Conclusions Both model and experiments show a significant impact of the liquid injection location. When the injection location is moved from the suction plenum to the start phase of compression, the model gives indicated isentropic efficiency increases from 5 to 50% while the experiments show an increase from 10 to 35%. The lower increase for the experiments most probably results from the smaller amount of injected liquid: 30 to 35 g sK1 in the experiments against 459 g sK1 in the model. The ideal location for liquid injection is during the start phase of compression. The labyrinth leakage flow is substantial and has a very large impact on the compressor performance. Both model and experiments show a significant effect of liquid injection on compressor performance. Also liquid injection from the high pressure side leads to significant improvement of the compressor performance. Acknowledgements This study was conducted as part of the NOVEM BSENEO Programme (BSE-NEO 0268.02.03.03.0002) and was partly financed by NOVEM.
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