Experimental and economic analysis with a novel ejector-based detection system for thermodynamic measurement of compressors

Applied Energy 261 (2020) 114395

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Experimental and economic analysis with a novel ejector-based detection system for thermodynamic measurement of compressors

T

⁎

Xiaoqiong Lia,1, Xiaoyan Wanga,1, Yufeng Zhanga, Lei Fangb, Na Denga, , Yan Zhanga, Zhendong Jinc, Xiaohui Yud, Sheng Yaoa a

Tianjin Key Lab. of Indoor Air Environmental Quality Control, School of Environmental Science and Engineering, Tianjin University, Tianjin 300072, PR China International Centre for Indoor Environment and Energy, Department of Civil Engineering, Technical University of Denmark, Lyngby 2800, Denmark c Shandong Chambroad Petrochemicals Co,. Ltd, Shandong Province, 256500, PR China d School of Energy and Environment Engineering, Hebei University of Technology, Tianjin 300401, PR China b

H I GH L IG H T S

ejector-based detection system was proposed. • AA novel method was proposed to calculate the ejector structure. • Thenumerical performance of the compressor was investigated. • The thermodynamic discharge temperature and COP of the virtual HTHP reached 184.8 °C and 3.94. • Economic comparison between ejector-based and HTHP detection system are evaluated. •

A R T I C LE I N FO

A B S T R A C T

Keywords: Ejector-based detection system High-temperature Compressor Virtual heat pump Economic analysis

The research of the high-temperature heat pump compressors above 100 °C has brought great challenges by its detection system, and in this study, a novel ejector-based detection system has been proposed to measure the thermodynamic performance of the heat pump compressors. The paper began with a numerical method to design the nozzle and diffuser of the ejector. Compared with the results of gas dynamics calculation, the deviations of nozzle and the diffuser were lower than 10.38% and 3.38%, respectively. Furthermore, the compressor performance was analyzed at the suction temperature in the range of 90–110 °C while the speed in the range of 2820–3060 r/min. The results showed that the compression ratio varied from 2.3 to 2.7, and the volumetric efficiency as well as isentropic efficiency was 90% and 70%, respectively. Assuming that the compressor was used in a heat pump system at the same operating condition, it was noteworthy that the highest discharge temperature was 184.8 °C and the coefficient of performance was more than 3.9. In an actual case, the initial investment cost of the ejector-based detection system is 3311 $, while that of the heat pump detection system is 59,795 $, the economical saving can reach to 94.5%. This is due to the exchanger has been omitted in ejectorbased detection system compared with the traditional heat pump detection system. All in all, the novel system was reliable and economical.

1. Introduction At present, industrial energy consumption occupies more than 30% from the total energy consumption in the energy consumption structure [1], while 70% of the total final energy demand in the industry sector is used for process heat [2], especially for the steam utilization and waste heat treatment. As we all known, the low-pressure steam of 1.0 MPa

mainly from the combustion of primary energy is to be extensively used in various industrial processes (chemical plant, food production, dyeing, etc.). However, the utilization efficiency of primary energy is very low [3] and more than half of the energy (waste heat) would be directly discharged into the environment through air coolers or water coolers in industrial processes [4]. High-temperature heat pump (HTHP) is a recognized energy-saving technology that can solve these

⁎

Corresponding author. E-mail address: [email protected] (N. Deng). 1 Co-first authors. https://doi.org/10.1016/j.apenergy.2019.114395 Received 18 September 2019; Received in revised form 25 November 2019; Accepted 14 December 2019 0306-2619/ © 2019 Elsevier Ltd. All rights reserved.

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Nomenclature A C COP CRF d dxi EEV f F GWP H, h i K L M n ODP P Pc Q S t,T Tc W

x isentropic index X average value of samples Z cost ($) α standard deviation ρ density (kg/m3) v specific volume (m3/kg) vs inspiratory flow (m3/r) ω velocity (m/s) ηs isentropic efficiency ηV volumetric efficiency ηP efficiency of pump τ time (s) φ maintenance factor 1,2,3,4,5,2a,3a,P,H,C,P* state point; section of ejector comp compressor cond condenser evap evaporator HX heat exchanger i, i + 1 micro element node in inlet k each equipment LMTD logarithmic mean temperature difference (°C) max maximum value min minimum value R reference value rated rated condition CI capital investment OM operation and maintenance cost

the area of heat exchanger (m2) state point of ejector diffuser coefficient of performance capital recovery factor diameter (mm) cross-sectional spacing (mm) electronic expansion valve cross-sectional area of the ejector (m2) correction factor of the logarithmic mean temperature difference global warming potential state point of ejector; enthalpy (kJ/kg); heating rate of boiler interest rate total heat transfer coefficient of heat exchanger (W/m2·K1) length (mm) mass flow rate (kg/s) speed (r/min); life time (year) ozone depletion potential pressure (MPa); state point of ejector inlet critical pressure (MPa) heating (kW) entropy (kJ/(kg·K)) temperature (°C) critical temperature (°C) power consumption (kW)

methods are all on the basis that the compressor is one part of the HTHP system, and the heat and cold source of HTHP system need to be equipped at the same time. The pressure-bearing and closed cooling water system should be adopted at the sink temperature over 100 °C [9]. The measuring results can be achieved with help of these methods, but they will lead a huge cost. The larger scale of the system is, the higher are the initial capital cost and operation cost of the heat and cold source. Therefore, it’s urgent to develop a new method to test the performance of the HTHP compressors with low cost. The ejectors have attracted more and more attentions due to their low cost and other excellent characters, such as simple structure, easy to be fabricated, stable and reliable operation [15]. The ejectors are maturely utilized on the study of the desalination applications [16], organic Rankine cycle system [17], vacuum device [18] and so on. In recent decades, the ejectors have also been proposed to improve the performance of the heat pump or refrigeration cycle. Chen et al. [19] proposed an integration system of an ejector with vapor-compression heat pump to enhance the heating performance for water heating application. The results showed that the maximum COP could be improved by 1.62–6.92% compared with the conventional heat pump. Sun et al. [20] presented an ejector heat exchanger based on an ejector heat pump and water-to-water heat exchanger. It improved the heating capacity of the existed primary heating network by approximately 43%. Besides, Li et al. [21] showed that the COP increased by 20% than that of the standard vapor-compression refrigeration cycle when the ejector was regarding as an expansion device in vapor-compression refrigeration cycle. The novelty of the present study is that a simple and economical ejector-based detection system for testing the performance of largescale high-temperature compressors is proposed. The new system replaces the evaporator, condenser as well as some equipment of the heat and cold source of the traditional HTHP system, which can reduce the area and volume of the system. With help of the new system, it can provide heat at the temperature of 184.8 °C, which can be extensively

problems and further promote the energy efficiency. As one of mature techniques, HTHP can recover low-temperature waste heat (especially below 100 °C) and elevate the low grade heat to high grade to replace the steam. In general, there are two types of vapor-compression HTHP products with the condensing temperature of 100–160 °C [5]: single-stage and multistage heat pumps. Recently, Chamoun et al. [6] designed and tested a new single-stage heat pump using water as refrigerant in the condensing temperature range of 130 °C–140 °C. Furthermore, a “BY” series refrigerants self-developed by Zhang’s group were used in HTHP, their experimental results showed that the condensing temperature of the single-stage HTHP (BY5) could reach 130 °C [7] and that of the cascade heat pump could reach to 142 °C (BY3A/R245fa, BY3B/ R245fa) [8] and 168 °C [9] (BY3A/BY6, BY3B/BY6), respectively. Meanwhile, the company of Ochsner [10] designed a HTHP with R134a/ÖKO1 as two-stage refrigerants, the sink temperature could vary from 95 °C up to 130 °C. Besides, Mateu-Royo et al. [11] theoretically evaluated the performance of five vapor-compression system configurations using low GWP refrigerant at temperatures of 110, 130 and 150 °C. As mentioned above, although the research of the HTHP has made great achievements, the higher temperature research still needs to be further explored in depth, while the associated technical barriers need to be addressed to expand the application scope of the HTHP in industries. The compressor is the key component of the vapor-compression HTHP. The performance of the HTHP can be reflected by that of the compressor [6]. It is well known that the compressors should be tested with performance first before putting into the market from the manufacture. The thermodynamic performance of the compressors is also very crucial on consideration of practical manufacturing except the compressors fault diagnosis [12], leakage detection [13] and related electrical testing. The methods of the thermodynamic performance testing mainly include two categories according to the standard [14]: refrigerant and the secondary fluid. Noteworthy, these detection 2

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experimental instrument. All data is recorded through the PID control panel. Detailed information of the measuring equipment is listed in Table 3.

used to the industrial processes. What else, a numerical method based on a Matlab-Refprop interface is utilized for calculating the nozzle and diffuser of the ejector in this work. The discharge temperature and compression ratio, volumetric efficiency and isentropic efficiency of the compressor at high temperature has been concerned and analyzed in this method, thus the feasibility of the ejector-based detection system is verified. It is worthy to be noted that the performance of the ejectorbased detection system is analyzed under the assumption that the compressor is used in a virtual heat pump system. Finally, an actual case has been discussed to illustrate the remarkable economic advantages of the ejector-based detection system compared with the traditional heat pump detection system. All in all, the ejected-based detection system proposed in this work has relatively high potential on commercial promotion.

2.3. Design of the ejector The reasonable design of the ejector is essential in order to ensure the normal operation of the detection system. As shown in Fig. 4, the ejector consists of nozzle, mixing section and diffuser [23].The performance of the ejector mainly depends on the geometric shape and the structure of the nozzle [15] and the mixing section [24]. In the past, the shape and structure of the ejector was mainly obtained by empirical formulas, which were used for certain specific fluids, not all fluids. Nowadays, the simulation software CFD [25] and some other numerical methods [26] are used to design the ejectors. In this paper, a numerical method is proposed to calculate the nozzle and diffuser of the ejector, which is based on the selected working fluid BY6. The numerical method improves the empirical formula calculation method of gas dynamics. This may be related to two aspects: (1) the geometric size of the ejector is obtained synchronously with the thermodynamic calculation; (2) the obtained thermodynamic calculation results are based on the reasonable change of working fluid state. The ejector is designed according to the experimental conditions in this paper. The pressure of the first fluid (primary fluid) inlet P, the second fluid (ejection fluid) inlet H and the mixed fluid outlet C of the ejector are 2.733 MPa, 0.34 MPa and 0.85 MPa, respectively. The mass flow rate of the primary fluid is 9040 kg/h, the entrainment ratio is equal to 0.17. According to the calculation in reference reported by Sokolov et al. [27], the maximum entrainment ratio can reach 0.185, which is higher than 0.17. Therefore, the design of the ejector is reasonable in theory. Thus, the nozzle is regarded as an example to introduce the model in the following Section.

2. System description 2.1. Refrigerant A near-azeotropic mixture refrigerant named BY6 with high critical temperature and low critical pressure was used as the working fluid of the ejector-based detection system. It was also used to design the twinscrew compressor in this paper. The parameters of BY6 was shown in Table 1. It should be noted that the ODP of BY6 is high. According to the Montreal Protocol, such refrigerant is only temporarily used to overcome the technical barriers encountered by various components of HTHP, such as the sealing and clearance problems of the screw compressor at high temperature. Moreover, the refrigerants with excellent thermodynamic performance and harmless to the environment are also being developed. All the specific parameters could be calculated through the software Refprop 9.1 [22] and the saturated parameters could be obtained from the reference reported by Li et al.[9]. 2.2. Description of the ejector detection system

2.3.1. Physical model of nozzle design For the sake of simplicity, the assumptions of the model are as follows:

Figs. 1 and 2 show the schematic diagram and P-h diagram of the ejector-based detection system, respectively. Fig. 3 shows the installation diagram of the system, the experimental research was carried and finished in a chemical plant in Shandong province, China. The ejector-based detection system mainly includes main cycle (solid line) and ejection cycle (dashed line), as shown in Fig. 1. The numbers in Figs. 1 and 2 represent different state points, and they are used in the formulas in the following Sections for modeling and analysis. The main cycle is composed of compressor, ejector, regulating valve 1 and a buffer tank, which is implemented to test the performance of the compressor. The steam coil is wrapped around the buffer tank. Initially the refrigerant in the buffer tank is heated up to 90 °C by the high-temperature steam (1.0 MPa). At the stable operation condition, the high-temperature steam is shut down and no steam will be provided to the system. The ejection cycle is composed of ejector, regulating valve 2, enclosed cooler and an electronic expansion valve (EEV). The EEV is used to regulate the mass flow of refrigerant. It is noteworthy that the ejector in the main cycle and the ejection cycle is the same equipment. The liquid storage tank and regulating valve 3 are only considered as standby system, which can provide sufficient refrigerant when the working fluid in the system is insufficient. Similar to the buffer tank, the steam coil is also wrapped around the liquid storage tank. The sight glasses are used to observe the refrigerant state in actual operation. All pipes in the system are packed with insulation cotton to reduce heat dissipation to the environment. Parameter specifications of main components of the system are listed in Table 2. Plug-in temperature sensor (PT100, ± 0.3 °C) and pressure sensor (0.2%F.S) are used to measure temperature and pressure of the each state point, respectively. The power meter and tachometer are installed on the electric control cabinet of the compressor. All sensors are calibrated before installation in order to reduce the error of the

(1) The whole flow processes were regarded as isentropic adiabatic. (2) The difference of viscous force between fluid and wall at each section was ignored. (3) There were no vortices and disturbances in all sections of the fluid. (4) The velocity on each section was assumed constant. Fig. 4 shows that the primary fluid enters the nozzle entrance (P) at initial velocity ωp, then the velocity increases to the sound velocity ωP* in the throat (P*). It continues to expand. At the exit of the nozzle, it reaches the supersonic speed ω1. Meanwhile, the pressure is slightly lower than the ejection pressure (PH). 2.3.2. Mathematical model of the nozzle design Using the finite element method of numerical calculation and taking a volume micro-element as a calculation unit, the nozzle can be calculated according to the flow chart, as shown in Fig. 5. The calculation methods of the ejector is based on a Matlab-Refprop interface. Matlab 2015b allows the dialogue between Refprop and Matlab. Specific steps are as follows: (1) the initial conditions are given in the software Matlab, which are the import parameters of primary fluid (temperature ΤP, pressure PP, mass flow rate M and the velocity ωP). The entropy S0, Table 1 The parameters of refrigerant BY6.

3

Substance

Molar mass (kg/ kmol)

Tc (°C)

Pc (MPa)

ODP

GWP (100 yr)

Safety

BY6

118.34

203.55

4.199

0.08

670

A1

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Fig. 1. Schematic diagram of the ejector-based detection system.

di + 1 = 2

fi + 1 (3)

π

In order to calculate the cross-sectional spacing dxi of the volume of the micro-element at each interface in the same temperature drop, the volume of each micro-element circular truncated cone should be calculated separately.

Mνi =

π d 2 d2 d ·d dx i × ⎛ i − 1 + i + i − 1 i ⎞ τ 3 4 4 4 ⎠ ⎝ ⎜

⎟

(4)

where τ represents the time when the fluid passes through the microelement circular truncated cone. The accumulation of section spacing is the length of nozzle, which can be calculated as follows: Fig. 2. P-h diagram of the ejector-based detection system.

i

L= enthalpy Η0 and specific volume v0 are calculated by invoking the software Refprop; (2) the step length is set to dt = −0.05 °C, ti +1 = ti + dt. The whole process is assumed as an isentropic process, that is, the constraint condition is Si+1 = S0; (3) each state parameter of node (i + 1) is calculated; (4) the value of state i + 1 is assigned to i + 2, which is iterated to the termination state in turn. When the pressure value of node i + 1 is the design value PH, the iteration is stopped. The state parameters, section size and horizontal length of the nozzle are obtained. The energy equation can be expressed as:

1 2 1 ωi + Hi = ωi + 12 + Hi + 1 2 2

0

M ρi + 1 ωi + 1

(5)

The numerical calculation method presented in this paper is applicable to all pure and mixed working fluids. The nozzle structure can be accurately designed according to the specific physical properties of working fluids and design conditions. 2.3.3. The overall structure of the ejector Currently, there is still no better calculation methods for the mixing section, while the existence of irreversible losses (collision loss, friction, and mixing loss). The empirical formula is adopted in this paper. According to the Ref. [27], the ejector is designed and the results are shown in Table 4. The ejector calculated by such empirical formula was considered reliable because it has been verified by experiments before [28]. The flow processes of the diffuser and nozzle are opposite and the mass flow rates are different. The diffuser can also be calculated by numerical method according to the condition of the ejector outlet, so its design procedure is similar to that of the nozzle. The ejector results of the numerical calculation and gas dynamics empirical method are

(1)

The sectional area and diameter of node i + 1 can be expressed by using Eqs. (2) and (3):

fi + 1 =

∑ dxi

(2) 4

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Fig. 3. Installation diagram of the experimental device.

temperature, compression ratio, volumetric efficiency and isentropic efficiency of the compressor are investigated when the compressor suction temperature is in the range of 90–110 °C and the speed is in the range of 2820–3060 r/min. All experimental data is the average values of multiple measurements after the system reached stability, with a maximum standard deviation below 1.2. The standard deviation can be

shown in Table 4. The results show that the maximum relative deviations of the nozzle and diffuser calculated by numerical calculation and gas dynamics empirical method are less than 10.38% and 3.38%, respectively. Therefore, the ejector designed by numerical method is reliable. Interestingly, the horizontal length of nozzle and diffuser has a great relationship with time when numerical calculation method is adopted. At the same thermodynamic condition, the different values of time may cause a great difference in the horizontal length of the nozzle and diffuser. Therefore, the length of the nozzle and diffuser can be calculated by empirical formula (the results are shown in Fig. 4).

n

∑ (Xi − X )

calculated by α =

i=1

n−1

, where, α is standard deviation, n is the

number of samples and X is the average value of samples. Fig. 6 shows the variations of the discharge temperature and compression ratio with the compressor suction temperature and speed. Compression ratio refers to the ratio of the discharge pressure to the suction pressure. As shown in Fig. 6(a), when the suction temperature increases, the discharge temperature of the compressor shows increase tendency first and then trend slows down gradually. This is due to the temperature of the rated design of the compressor is 110 °C. At the same suction temperature, the discharge temperature of the compressor increases upon an increase in the compressor speed. Moreover, the content in the dotted frame in Fig. 6(a) shows that the saturation discharge temperature of the compressor can reach 160 °C at the corresponding suction temperature and compressor speed, which has a wide range of applications in industries. When the suction temperature is equal to 110 °C and the speed of the compressor is 3060 r/min, the maximum discharge temperature never exceeds 184.8 °C, which is lower than the alarm temperature (190 °C) of the compressor. It is noteworthy that the temperature of 184.8 °C is one of the highest heating temperature in the vapor-compression HTHP system [9]. The stable operation of the compressors at high-temperature plays a very important role in the popularization of the HTHP in industries. Fig. 6(b) shows that the compression ratio of the compressor decreases slightly at the same speed upon an increase in the suction temperature. The reason of this phenomenon can be explained by the fact that the increasing trend of the discharge temperature and of the discharge pressure gradually slows down upon an increase in the

2.4. Test conditions The feasibility and reliability of the ejector-based detection system are investigated and analyzed by measuring the thermodynamic performance of the compressors under different suction saturation temperature and speed. The suction saturation temperature ranges from 90 to 110 °C, the increment step is 5 °C. The suction pressure corresponds to 0.54–0.85 MPa. The speed of the compressor is in the range of 2820–3060 r/min and the increment step is 60 r/min. The test conditions of the experimental system are shown in Table 5. 3. Experimental results and analysis of the ejector-based detection system 3.1. The compressor section The evaluation of the compressor performance is the basis of whether the ejector-based detection system is reliable or not. The hightemperature open oil-free twin-screw compressor is used as the testing equipment in this paper. Ignoring the heat dissipation from the compressor to the environment, it is assumed that the fluid is in a steady state in the system. In this Section, the variations of the discharge 5

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suction temperature. Furthermore, the increase rate of discharge pressure is lower than that of the suction pressure. At the same suction temperature, the compression ratio of the compressor increases upon an increase in the compressor speed. This is due to the leakage of the compressor will be reduced and the mass flow rate increases with the increase of the compressor speed. The results are consistent with the change trend of the discharge temperature with the speed of the compressor. Fig. 6(b) shows that the compression ratio is in the range of 2.3–2.7, which is reasonable for the single stage compression ratio of the open oil-free twin-screw compressor, corresponding to the results (1.5–3.5) reported by Xing et al. [29]. The volumetric efficiency and isentropic efficiency of the compressor can be expressed as Eqs. (6) and (7) [30]:

2 1 3

1

1 1

/ / Anshan yuan He Zheng Tong automatic control instrument Co., Ltd. Wuxi Dingbang heat exchange equipment Co., Ltd.

Shandong Chambroad Petrochemicals Co,. Ltd. Shandong Chambroad Petrochemicals Co,. Ltd.

GP3 × 3-4-42-1.6S-23.4/DR

x−1

T1 ⎡ (P2 P1 ) x − 1⎤ ⎦ ηV = ⎣ T2 − T1

ηs =

(6)

h2s − h1 h2 − h1

(7)

where x = 1.2, h2s is the outlet enthalpy of the compressor in isentropic compression. Fig. 7 shows that the volumetric efficiency and isentropic efficiency remain basically unchanged upon an increase in suction temperature. As can be seen from Fig. 7(a), at the same suction temperature, the volumetric efficiency of the compressor will increase slightly with the increase of the compressor speed. The trend is approximately in line with the compression ratio. However, Fig. 7(b) shows that the compressor speed has a small effect on the isentropic efficiency. The volumetric efficiency of the compressor in this system is about 90% and the relative deviation is less than 0.03. It is within a reasonable range of 75–95% of the volumetric efficiency proposed in Ref. [29]. Similarly, the isentropic efficiency of the compressor is about 70% and the relative deviation is below 0.01. The isentropic efficiency of screw compressor provided in Ref. [29] is 65–75%. It can be deduced that the volumetric efficiency and isentropic efficiency of the compressor in this system are reasonable. Not only the design condition of the high-temperature twin-screw compressor can be realized, but also the compression ratio, volumetric efficiency and isentropic efficiency of the compressor can be measured in a reliable range by the ejector-based detection system. Therefore, the ejector-based detection system can be used to detect high-temperature twin-screw compressor. 3.2. The virtual heat pump system section

V-101 V-102

The qualified compressor will eventually be used in the HTHP system. Moreover, the test condition of the ejector-based detection system is the operating condition of the compressor in the HTHP. Therefore, the data obtained in the ejector-based detection system can be used to calculate the performance of the HTHP, which is defined as a virtual HTHP. The heating capacity and COP are important for the HTHP. When the HTHP system runs steadily, the mass flow of the compressor can be calculated as follows:

m = ηV ρcomp, in V = ηV ρcomp, in

vs ·n nrated

(8)

The refrigerant at the compressor outlet is condensed under a constant pressure, then turns into the saturated liquid or has 3–5 °C supercooling degree at the outlet of the condenser in a traditional HTHP system. However, in the ejector-based detection system, the temperature and pressure at the inlet and outlet of the condenser are not measured directly, due to that there is no condenser in such system. In the virtual HTHP system, it is assumed that the refrigerant is condensed to a saturated liquid. The calculation of the virtual heating capacity can be expressed as follows:

Buffer tank Liquid storage tank

Enclosed cooler

YE2-100L2-4 AL-4 DT-26 K

1 Hengshui electric Limited by Share Ltd YVFE2-280S-2

Variable Frequency Adjusting Speed 3-phase Induction Motor Compressor lubrication oil pump Oil cooler Electronic expansion valve (EEV)

1 KLY127.5–100 Open twin-screw compressor

Dandong Colossus

The material of the nozzle is stainless steel and the shell is cast steel. Specific parameters refer to Section 2.3 The twin-screw compressor's suction volume is 4 m3/min and its rotate speed is 3000 r/min, rated power is 120 kW Rated power 120 kW, rotate speed is 3000 r/min, voltage is 380 V, triangle connection method The rotate power is 3 kW, the frequency is 50 Hz, and the rotate speed is 1440 r/min The cooling area is 4 m2, the design pressure is 3 MPa, and the design temperature is 110 °C The temperature range of the valve is −29 to 200 °C, nominal pressure is 2.5 MPa, input signal is 4–20 mA, voltage AC220V/50 Hz Fan diameter is 1000 mm, four blades, motor rated power is 5.5 kW, speed is 970 r/min, speed adjustable; air volume is 4 * 104 m3/h The main size is DN800 * 1600. The silk fishing net size is DN700 * 100 The main size is DN600 * 1200 1 Tianjin Bainiao Machinery Manufacturing Co., Ltd. Ejector

Customized

Quality Manufacturer Type Equipment

Table 2 Parameter specifications of main components of the system.

Specification

X. Li, et al.

6

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Table 3 Specifications and uncertainties of the test devices. Name

Type

Uncertainty

Range

Manufacture

Pt100 Pressure sensor (compressor suction) Pressure sensor (compressor discharge) Data acquisition

H-WZYK 3051CG4A22A1AM5B4DFI3Q4HR5 3051CG4A22A1AM5B4DFI3Q4HR5 PID collection

± 0.3 °C 0.2%F.S 0.2%F.S /

0–300 °C 0–1.6 MPa 0–4.0 MPa /

Chongqing Chuanyi instrument No.17 factory Co.,Ltd Beijing Far East instrument Co., LTD Beijing Far East instrument Co., LTD Dandong Colossus

Fig. 4. The ejector structure and size. Table 4 Calculation results of ejector structure based on numerical calculation and gas dynamics empirical method.

Numerical calculation Gas dynamics empirical method Relative deviation

mP kg/h

dp mm

dP* mm

d1 mm

d4 mm

d3 mm

dc mm

9040 9040

43.70 43.70

17.07 15.60

26.50 23.75

34.85 33.97

33.09 31.97

65.51 63.58

/

0.000

0.086

0.104

0.025

0.034

0.029

Table 5 Test conditions of the experimental system.

(9)

where h2′ is the enthalpy of saturated liquid corresponding to the refrigerant pressure at the outlet of the compressor. The COP of the virtual HTHP system can be expressed as follows:

COP = Q Win

Unit

Range

Step

Suction saturation temperature Suction pressure Discharge pressure Speed of the compressor

°C MPa (A) MPa (A) r/min

90–110 0.54–0.85 1.29–2.11 2820–3060

5 / / 60

Fig. 8 shows the variations of the heating capacity, heating capacity of unit mass, input power and COP with the suction temperature and compressor speed. The left and right arrows in Fig. 8(a) indicate that the value of heating capacity and heating capacity of unit mass, respectively. As can be seen from Fig. 8(a), the unit mass heating capacity of BY6 decreases upon an increase in the suction temperature at the same speed. This is due to that the higher the discharge temperature is, the lower is the specific latent heat of BY6. The reverse change of the heating capacity and the heating capacity of unit mass is due to the fact that the average increase rate of mass flow of refrigerant is 11.6%, which is larger than the decrease rate of unit mass heating capacity by 3.6%. Finally, the average increase rate of heating capacity of the system is 7.6% upon an increase of the suction temperature by 5 °C. At the same suction temperature, the unit mass heating capacity decreases, while the heating capacity rises with the increase of the compressor speed. This is due to that the increase rate of mass flow is 2.1% and the reduction rate of unit mass heating capacity is 1% for every 60 r/min of the compressor speed. Noteworthy, the heating capacity of the system is more than 220 kW. Fig. 8(b) indicates that the input power increases upon an increase in the suction temperature. This is because the increase of the suction density of the compressor leads to the increase of the mass flow. In addition, when the speed of the compressor increases at the same

Fig. 5. Flowchart of nozzle calculation for ejector.

Q = m (h2 - h2′)

Parameter

(10)

where Win is the input power of the compressor. 7

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Fig. 6. Variations of the discharge temperature and compression ratio with the compressor suction temperature and speed.

suction temperature, the suction volume flow rate will be increased, thus, the mass flow rate and the input power will be increased in correspondingly. The average increase rate of input power is lower than 14% with the increase of the suction temperature of 5 °C and it is lower than 7% with the increase of the compressor speed of 60 r/min. Fig. 8(c) shows that the COP decreases with the increase of the suction temperature and the compressor speed. This is due to when the compressor speed increases, the temperature lift of the inlet and outlet of the compressor gradually increases. In addition, The COP attains its minimum value of 3.94 at the evaporation temperature of 110 °C and the compressor speed of 3060 r/min. The minimum COP is still higher than 2.5, which illustrates that the HTHP system has economic advantages when using such compressor.

because of the low investment costs compared with large equipment. The design parameters of this case and the correlation coefficients are listed in Table 6. The capital investment cost of each component is calculated based on the following cost functions. It should be pointed out that these cost calculations are used only for rough estimates and do not provide the current price of the equipment to be purchased. For traditional HTHP system, the cost of the evaporator and condenser can be calculated as follows [31]:

A ZHX = ZR . HX ⎛ HX ⎞ ⎝ AR ⎠ ⎜

0.6

⎟

(11)

where Z denotes the total cost. The subscript R is the reference state point, the specific value can be found in Table 6. A is the heat transfer area of the evaporator and the condenser, which can be expressed as follows:

4. Economic benefit comparison between ejector-based and heat pump detection system

A=

In this work, the economic benefit between the detection system of the ejector-based and traditional heat pump with heat and cold source has been compared. The simplified diagram of the two detection systems is shown in Fig. 9. The parts marked in red in Fig. 9 are different components in both systems. The economic analysis is mainly concentrated on these components, regardless of the same components. Economic analysis of pipeline and the control systems are neglected

Qevap

cond

KF ΔTLMTD

(12)

where K and F refer to the heat transfer coefficient and the correction factor of the logarithmic mean temperature difference, respectively. The logarithmic mean temperature difference is ΔTLMTD = (ΔTmax − ΔTmin ) ln(ΔTmax ΔTmin ) , which is equal to 5 °C in this paper.

Fig. 7. Variations of volumetric efficiency and isentropic efficiency with the compressor suction temperature and speed. 8

Applied Energy 261 (2020) 114395

X. Li, et al.

Fig. 8. Variations of (a) heating capacity, heating capacity of unit mass, (b) input power and (c) COP with the compressor suction temperature and speed.

Enclosed cooler

EEV

Water tank

Enclosed cooler

Pump

Condenser

Ejector

Compressor EEV

Compressor

Buffer tank Evaporator

Buffer tank

Boiler Pump

(a) Ejector-based detection system

(b) Heat pump detection system

Fig. 9. Simplified diagram of the detection system. 9

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In this case, according to the above economic model, the economic comparison results between the ejector-based and heat pump detection system are listed in Table 7. Table 7 clearly shows that the initial investment cost of the ejectorbased detection system is 3311 $, while the initial investment cost of the heat pump detection system is 59,795 $. It is obvious that the initial system invests 94.5% more than the ejector-based detection system. The most expensive equipment in heat pump detection system is the heat exchanger, accounting for 59.3% of the total investment. Interestingly, the heat exchanger has been omitted in the ejector-based detection system, which can greatly save the cost. Considering the initial investment, operation and maintenance cost of the system, the average annual cost of the ejector-based and heat pump detection system are 561 $/year and 10,126 $/year, respectively. In conclusion, for one thing, the ejector-based detection system can not only realize the function to detect the performance of the compressors, but also greatly reduce the cost compared with the traditional heat pump detection system. For another, several large-scale devices have been omitted in the ejector-based detection system, which can reduce the area and volume of the system. Such detection system can be considered by enterprises in the choice of a new compressor detection system.

Table 6 Design parameters of the case. Parameters

Unit

Value

Ref.

Evaporation temperature of refrigerant Condensing temperature of refrigerant The speed of the compressor, n Heating capacity of condenser, Qcond Heating absorption of evaporator, Qevap Overall heat transfer coefficient of the evaporator and condenser, K Correction factor of the logarithmic mean temperature difference, F The logarithmic mean temperature difference (ΔTLMTD ) The reference cost of evaporator and condenser, ZR Reference area of heat exchanger, AR The reference cost of pump, ZR.P The reference power consumption of Pump, WR.P Pump efficiency, ηP Heating Rate of Boiler, H Annual interest rate, i The component total operating lifetime, n Maintenance factor, φ

°C °C r/min kW kW W/(m2·K)

110 160 3000 312 234.4 1200/750

/ / / / / [35]

/

0.93

[36]

°C

5

[37]

$

[31]

m2 $ kW

16,000/ 8000 100 2100 10

[31] [31] [31]

% 1/kW % year /

90 83,700 15 20 0.06

[31] [33] [34] [34] [34]

5. Conclusion The investment cost of the water pump [31], enclosed cooler [32] and hot water boiler [33] can be expressed as follows, respectively:

W ZPump = ZR . P ⎛ P ⎞ W ⎝ R.P ⎠ ⎜

ZCooler =

0.26

⎟

⎛⎜ 1 − ηP ⎞⎟ ⎝ ηP ⎠

In this paper, a novel ejector-based detection system for measuring the thermodynamic performance of high-temperature compressors was presented. The main conclusions were drawn and summarized as follows:

0.5

(13)

0.76 629.05WCooler

(14)

Zboiler = 205H−0.13Q

(1) A numerical method was proposed for calculating the structure of nozzle and diffuser of the ejector, which is suitable for pure and mixed fluids. It guarantees the accuracy of the detection system performance using the fluids properties to design solid structures. Compared with the results of gas dynamic calculation, the maximum deviations of nozzle and diffuser are lower than 10.38% and 3.38%, respectively. (2) In the ejector-based detection system, the compression ratio is in the range of 2.3–2.7. The average volumetric and isentropic efficiency of the compressor are 90% and 70%, respectively, which are all within a reasonable range. The feasibility of the ejector-based detection system has been verified. (3) When the evaporator temperature is higher than 105 °C and the speed of the compressor is more than 3000 r/min, the maximum discharge temperature is 184.8 °C and corresponding to the saturation temperature reaches 160 °C. The heating capacity of the virtual high temperature heat pump is more than 220 kW and coefficient of performance is more than 3.94. (4) The initial investment of the ejector-based and traditional heat pump detection system are 3311 $ and 59,795 $, the average annual cost of them are 561 $/year and 10,126 $/year, respectively.

(15)

where W is the power consumption, ηP is the pump efficiency. H and Q are the heating rate and heat of boiler, respectively. The total investment of each component can be expressed as [34]: ·

·

·

k k Z k = ZCI + ZOM

(16) ·

k where k corresponds to each equipment in the system. ZCI and

·

k ZOM represent the annual capital investment and annual operation and .

·

·

maintenance cost, respectively. ZCI = CRF ·Zk and ZOM = φ ·ZCI . Capital recovery factor (CRF) for interest rate (i) and life time (n) can be exi (1 + i)n pressed as CRF = (1 + i)n − 1 . φ is the maintenance factor. The total annual capital investment of the system is the sum of the investment of each component, which can be expressed as follows: ·

·

Ztotal =

∑ Zk

(17)

For ejector-based detection system, the price of the ejector is provided by the manufacture. The difference between the enclosed cooler in two detection systems is the difference in power consumption. Table 7 Economic comparison between ejector-based and heat pump detection system. Component

Evaporator Condenser Enclosed cooler Ejector Water pump in heat sink Water pump in heat source Hot water boiler Overall system

Ejector-based detection system

Heat pump detection system .

.

.

.

Capital investment cost ($)

ZCI ($/year)

ZOM ($/year)

Capital investment cost ($)

ZCI ($/year)

ZOM ($/year)

0 0 2298 1013 0 0 0 3311

0 0 367 162 0 0 0 529

0 0 22 10 0 0 0 32

26,923 8553 5658 0 931 821 16,908 59,795

4301 1366 904 0 149 131 2701 9553

258 82 54 0 9 8 162 573

10

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X. Li, et al.

Compared with the traditional heat pump detection system, the cost of the ejector-based detection system can be reduced by 94.5%, which has a great economic benefit. For large-scale high temperature compressors’ detection, such system is worth considering by enterprises.

[10] [11]

[12]

CRediT authorship contribution statement

[13]

Xiaoqiong Li: Conceptualization, Data curation, Formal analysis, Methodology, Software, Validation, Writing - original draft, Writing review & editing. Xiaoyan Wang: Conceptualization, Data curation, Formal analysis, Methodology, Validation, Writing - original draft, Writing - review & editing. Yufeng Zhang: Conceptualization, Data curation, Funding acquisition, Methodology, Resources, Software, Supervision, Writing - review & editing. Lei Fang: Conceptualization, Formal analysis, Validation, Writing - review & editing. Na Deng: Conceptualization, Methodology, Software, Supervision, Visualization, Writing - review & editing. Yan Zhang: Data curation, Investigation. Zhendong Jin: Funding acquisition, Resources, Validation. Xiaohui Yu: Conceptualization, Writing - review & editing. Sheng Yao: Software.

[14] [15] [16] [17]

[18] [19]

[20] [21]

Declaration of Competing Interest

[22] [23]

The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper.

[24] [25]

Acknowledgements [26]

The authors gratefully thanks the support provided by the 973 National Basic Research Program of China (No. 2015CB251403). And the authors also express their thanks to Shandong Chambroad Petrochemicals Co,. Lt, especially those who participated in the experiment: Jipeng Hao, Kai Zhang, Xin Zhong and Liehui Zhao.

[27] [28]

[29]

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[30]

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