energies

Article Innovative Turbine Intake Air Cooling Systems and Their Rational Designing

Andrii Radchenko 1, Eugeniy Trushliakov 1, Krzysztof Kosowski 2, Dariusz Mikielewicz 2,* and Mykola Radchenko 1

1 Department of Ship Electroenergetic Systems, Admiral Makarov National University of Shipbuilding, Heroes of Ukraine Avenue 9, 54000 Mykolayiv, Ukraine; [email protected] (A.R.); [email protected] (E.T.); [email protected] (M.R.) 2 Faculty of Mechanical Engineering, Gda´nskUniversity of Technology, 80-233 Gda´nsk,Poland; [email protected] * Correspondence: [email protected]



Received: 15 September 2020; Accepted: 20 November 2020; Published: 25 November 2020


Abstract: The efficiency of cooling ambient air at the inlet of gas turbines in temperate climatic conditions was analyzed and reserves for its enhancing through deep cooling were revealed. A method of logical analysis of the actual operation efficiency of turbine intake air cooling systems in real varying environment, supplemented by the simplest numerical simulation was used to synthesize new solutions. As a result, a novel trend in engine intake air cooling to 7 or 10 ◦C in temperate climatic conditions by two-stage cooling in chillers of combined type, providing an annual fuel saving of practically 50%, surpasses its value gained due to traditional air cooling to about 15 ◦C in absorption lithium-bromide chiller of a simple cycle, and is proposed. On analyzing the actual efficiency of turbine intake air cooling system, the current changes in thermal loads on the system in response to varying ambient air parameters were taken into account and annual fuel reduction was considered to be a primary criterion, as an example. The improved methodology of the engine intake air cooling system designing based on the annual effect due to cooling was developed. It involves determining the optimal value of cooling capacity, providing the minimum system sizes at maximum rate of annual effect increment, and its rational value, providing a close to maximum annual effect without system oversizing at the second maximum rate of annual effect increment within the range beyond the first maximum rate. The rational value of design cooling capacity provides practically the maximum annual fuel saving but with the sizes of cooling systems reduced by 15 to 20% due to the correspondingly reduced design cooling capacity of the systems as compared with their values defined by traditional designing focused to cover current peaked short-term thermal loads. The optimal value of cooling capacity providing the minimum sizes of cooling system is very reasonable for applying the energy saving technologies, for instance, based on the thermal storage with accumulating excessive (not consumed) cooling capacities at lowered current thermal loads to cover the peak loads. The application of developed methodology enables revealing the thermal potential for enhancing the efficiency of any combustion engine (gas turbines and engines, internal combustion engines, etc.).

Keywords: combustion engine; intake air; cooling capacity; chiller; current thermal load; annual fuel reduction

1. Introduction Fuel efficiency of combustion engines (gas turbines [1–3], diesel engines [4] and gas engines [5–7]) falls with raising an ambient air temperature at their intake. Gas turbines (GT) are especially sensitive to intake conditions [8–10]: a specific fuel consumption increases by 0.4 to 1.0 g/(kWh) for every 1K rise

Energies 2020, 13, 6201; doi:10.3390/en13236201 www.mdpi.com/journal/energies Energies 2020, 13, 6201 2 of 22 in intake air temperature. A lot of ambient air cooling technologies were developed to enhance engine output [11–13]. The cheapest and most widespread in hot climatic conditions is a contact method by evaporation of water sprayed into air stream at the intake of GT compressor [14] and a novel contact cooling according to Maisotsenko cycle [15,16], but their efficiency is limited by the magnitudes of ambient air wet bulb temperature and they are not so effective in temperate climatic conditions. Meantime the GT efficiencies are typically of 30 to 35% and a major part of the fuel energy consumed is dissipated to atmosphere as exhaust gases. Therefore, the use of exhaust heat from GT is a very effective trend in enhancing their overall efficiency [17,18]. Issuing from this, turbine intake air cooling (TIAC) in exhaust heat recovery chillers is a very perspective approach to improving the fuel efficiency of GT at increased intake ambient air temperatures [19–21]. Thus, cooling intake air by converting the heat of exhaust gas in chillers provides increasing the efficiency of GT with corresponding fuel saving at raised ambient air temperatures [22–24]. The absorption lithium-bromide chillers (ACh) are the most widely used in a hot climate [25–27]. They are characterized by a high efficiency: their coefficients of performance COP is 0.7 to 0.8 (COP is ratio of generated cooling capacity Q0, kW, to the heat consumed Qh, kW), environmentally friendly, and noise free [28–30]. The same advantages and ability to use a low potential heat with a bit less COP are peculiar for adsorption chillers [31,32]. The refrigeration vapor compression chillers require mechanical or electrical driving [33], while absorption and adsorption chillers use only waste heat. Some authors have investigated the use of ACh to improve the efficiency of contact methods of cooling, in particular by cooling the water injected into the air stream; however, the most common way of cooling air at the inlet of the GT is the application of air coolers fed by ACh cooling capacity. Many publications focused for improving ambient air processing by intensification of heat transfer processes in air coolers-evaporators [34] and condensers [35–37] of refrigeration chillers, in particular, through refrigerant recirculation to intensify heat transfer, increase heat flux and reduce temperature difference in intake air coolers as result [38,39], waste heat recovery technics for combined cooling, heating and power (CCHP), so-called trigeneration or integrated energy systems [40,41]. Some of principal technical innovations and methodological approaches in heat recovery: jet technologies [42–44], deep exhaust heat use [45–47] for increasing the available waste heat potential to be converted into refrigeration and others were developed for TIAC or might be successfully applied in TIAC to match current cooling demands [48–50], in particular, two-stage air cooling [51]. The intake air temperature depression ∆t = t ta , influencing the efficiency of TIAC, depends amb − 2 on the ambient air temperature tamb and a temperature ta2 of air cooled in chiller at the inlet of GT, which depends on the temperature of a coolant, i.e., on the type of chiller. In ACh of a simple cycle with high coefficients of performance COP is 0.7 to 0.8 and a temperature of chilled water tw of about 7 ◦C an intake air can be cooled only to ta2 = 15 ◦C. In the most simple in design and cheap refrigeration ejector chillers (ECh) that use refrigerants as low boiling coolant, the intake air can be cooled to lower temperatures ta2 is 7 to 10 ◦C at the temperatures of boiling refrigerant t0 ranging from 1 to 5 ◦C, but with low COP from 0.2 to 0.3 that requires raised amounts of exhaust gas heat. It is quite reasonable to apply a highly efficient ACh as a high-temperature stage of intake air cooling from current ambient air temperatures tamb to ta2 = 15 ◦C and a less efficient ECh as a low-temperature stage for further subcooling air lower than 15 ◦C. The application of such hybrid intake air coolers is especially expedient for operation of GT in temperate climatic conditions. They are able to provide deep intake air cooling with corresponding increase of TIAC operation duration and much more annual effect as a result: fuel reduction or power output increase, compared to traditional cooling in ACh. Such innovative two-stage ambient air cooling systems consist of chillers that convert the heat of exhaust gas into refrigeration in the form of coolant, feeding turbine intake air coolers, and could find effective application in different types of combustion engines: gas turbines [52,53], gas engines for combustion of biogas [54,55], stationary and marine power plants [56,57]. Energies 2020, 13, 6201 3 of 22

Many modern methods of analyses [58–60] are aimed to increase the effect gained due to ambient air processing, in particular, to reduce energy and fuel consumption. Some of such methods include calculation of cooling degree-hours (CDH) [14,61] and modified versions [62–64] to match current cooling duties according to actual climatic conditions as well as thermal demand management (TDM) [65,66] based on different criteria [67,68] to save energy in building conditioning, engine intake air cooling and combined cooling, heat and power. Cooling degree-hour (CDH) calculations are widely used for evaluating the efficiency of application of ambient air cooling in site climatic conditions in power generation—combustion engine intake air cooling [63,64], as well as in air conditioning [62]. So as the performance efficiency of TIAC systems needs to be analyzed across the full range of operating conditions, the CDH number is used to P determine the total amount CDHs in a particular climate over a considered period. The use of TIAC potential in terms of CDH provides easy calculation of fuel reduction and engine power output augmentation for any considered period. For this purpose a climatic characteristic of GT as dependence of specific fuel consumption be, g/(kWh), and GT power output Pe, kW, on intake air temperature are used. The ambient air temperature and relative humidity distribution during year or any cooling period is very important input data for energy analyses and a design cooling load determination. The studies on the input ambient data when evaluating gas turbine inlet cooling are presented in [62–64]. Although many researchers consider the cumulative CDH profile along with time elapsed [62], only a few studies focus on analyzing the behavior of yearly cumulative cooling profiles in dependence on loading to determine a design cooling load [63,65]. All the typical methods, based on summarizing the number of CDH, issue from the assumption of a design cooling capacity as a sufficient to provide maximum cooling needs over the full range of yearly operating conditions [63]. Such an approach may lead to considerable oversizing of the chillers and the TIAC systems in the whole that requires to solve the problem of determining the correct design cooling load excluding oversizing, as it was shown in [69–71]. A design cooling capacity Q0 of TIAC system, on the one hand, has to cover intake air cooling, and needs to do so during as long a time of GT operation during a year as possible, which provides the greatest annual fuel reduction. On the other hand, a design cooling capacity Q0 should not be overestimated so that for the most part of the year a TIAC system would be able to operate at a high load level close to a design value. Otherwise low efficiency of TIAC system operation takes place, and on the contrary, when underestimating Q0—insufficient turbine intake air cooling at high ambient temperatures. Thus, the fuel efficiency of GT in temperate climatic conditions firstly might be considerably enhanced due to deep intake air cooling that needs the application of combined chillers and rational designing of TIAC systems to guarantee a close to maximum annual fuel saving, but without system overloading, as an alternative to their traditional designing to cover the maximum loading with inevitable system oversizing. The goal of this investigation is to increase the fuel efficiency of GT by deep intake air cooling in combined absorption ejector TIAC systems with ACh as a high-temperature and ECh as a low-temperature cooling stages and through rational designing of the proposed systems.

2. Materials and Methods A method of logical analysis [72,73] of the actual operation efficiency of TIAC systems in real varying environment, supplemented by the simplest numerical simulation that enables its easy application in designing practice of any ambient air cooling system, is behind this research and the design methodology proposed. The real input data on site actual climatic conditions (ambient air temperature tamb and relative humidity ϕamb) were taken by using the well-known programs, for instance “meteomanz” [74]. Energies 2020, 13, 6201 4 of 22

To avoid errors caused by the approximation of the current changeable effect gained due to cooling (fuel reduction, power output increase, etc.) at varying loading of TIAC systems in response to actual climatic conditions, the fluctuations of the last are considered through summarizing the current values of the effect and its evaluated by different rate of its annual value increment. Thus, the novelty of the research methodological approach to analysis of the TIAC system operation efficiency consists of summation of the current changeable effect (fuel reduction, etc.) due to cooling air over the year as an annular effect (annual fuel saving). The rate of its increment according to cooling capacity is used as an indicator to efficient realization of cooling capacity generated. To realize the methodological approach mentioned above the improved methodology of TIAC system designing based on the annual effect is developed. It involves additional stages to determine an optimal value of cooling capacity, providing the minimum system sizes at maximum rate of annual effect increment, and its rational value, providing a close to maximum annual effect without TIAC system oversizing at the second maximum rate of annual effect increment within the range beyond the first maximum rate. The tradition methods of TIAC system designing is based on the approximation of varying current values of the effect in the form of cumulative effect characteristic built by the samples of corresponding current values of effect in dependence on loading. With this a design cooling load is selected about 20% higher than its value corresponding to the maximum of cumulative characteristic [65]. The most widespread methods of TIAC system designing are intended to provide the maximum values of current effect during a year or annual effect due to air cooling [63] and leads to overestimating design cooling capacity and TIAC system oversizing as a result. A proposed designing methodology based on the approach of two maximum rates of annual effect increment is very quantitative and simple for realization in different types of energetic plants including combustion engines and air conditioning systems.

2.1. General Assumptions and Hypothesis The assumptions adopted for the analysis of the operation efficiency of TIAC systems and their rational designing are as follows:

The lowest temperature of air cooled in ACh of a simple cycle is assumed to be ta = 15 C and • 2 ◦ limited by minimum temperature difference of 8 ◦C between cooled air and chilled water leaving ACh at tw = 7 ◦C (water at the inlet of air cooler): ta2 = tw + 8 ◦C. In the case of using a refrigerant as a coolant in the air cooler the temperature difference between • heat exchanging fluids is lower, 4 or 5 ◦C, falling in dependence of the refrigerant boiling point in air cooler within t0 of 2 to 5 ◦C. This leads to lower values of minimum temperature ta2 of air cooled in refrigerant chiller, for instance, 7 to 10 ◦C in ECh: ta2 = t0 + (4 or 5) ◦C. The last assumption regulates a joint operation of ACh and ECh in a two stage cooling air mode: • to 15 ◦C in ACh and further subcooling to 10 ◦C in ECh of a stage absorption-ejector chiller (SAECh) and in cascade mode at lowered ambient air temperatures tamb when ambient air temperature drops to tamb = 15 ◦C the excessive (not consumed) cooling capacity of ACh is used for cooling the condenser of ECh as a low cascade of cascade absorption-ejector chiller (CAECh). As well as proposed absorption-ejector TIAC systems are the advanced versions of typical basic • absorption TIAC system, the economic comparison with the last might be done taking into account only the cost of extra heat exchangers of ECh (refrigerant evaporator-air cooler, refrigerant condenser, and ejector) with unchanged maintenance cost, personal, etc. Because of fluctuations in cost of heat exchangers of different manufacturers and a fuel especially the economic analysis is to be conducted for the concrete case. Thus, the considered method of designing focuses to provide just initial basic data as rational technical characteristics for further complicated detailed economic analysis. Energies 2020, 13, 6201 5 of 22

The hypothesis accepted to prove novel approaches to the principles of proposed innovative TIAC system operation with account of actual climatic conditions are the following. The fluctuations of current thermal loads on TIAC systems, caused by variable actual ambient air parameters, should be covered by the chiller with high coefficients of performance COP, i.e., by ACh with COP from 0.7 to 0.8 that provide cooling ambient air to 15 ◦C. The further subcooling of air down to 10 ◦C and lower is conducted within a comparatively stable range of loading and can be covered by less efficient ECh (COP is from 0.2 to 0.3) and more sensitive to load changes. In temperate climatic conditions, when cooling load on ACh falls, the effect in reduced specific fuel consumption might be still considerably enlarged due to deeper cooling ambient air at the intake of GT to the temperatures from 7 to 10 ◦C, compared with its traditional cooling to 15 ◦C in ACh of a simple cycle. According to this hypothesis a deep GT intake air cooling is considered to be a perspective trend of enhancing turbine efficiency. To provide deep cooling the absorption-ejector TIAC system operates as a stage system to conduct two-stage cooling air consequently to the temperature of 15 ◦C by chilled water from ACh and further subcooling to about 10 ◦C by boiling refrigerant from ECh. At lowered ambient air temperatures, and thermal load on ACh accordingly, the excessive (not consumed) cooling capacity of ACh is used for cooling the condenser of ECh as the low cascade of cascade absorption-ejector chiller (CAECh) to provide deeper air cooling as compared with two-stage cooling air. The following hypothesis to prove the novel approaches to designing of TIAC systems by the proposed method of determining a rational design cooling capacity are accepted. P The annual fuel reduction B is used as a primary criterion. To avoid the errors caused by the approximation of the current changeable effect gained due to cooling (fuel reduction, power output increase, etc.) at varying loading of TIAC systems in response to actual climatic conditions, the fluctuations of the last are considered through summarizing the current values of the effect and are evaluated by different rate of its annual value increment. The summation of current values of changeable effect, gained due to air cooling at varying loading of TIAC system, over the considered period (year), allows considering their changes by the rate of its annual value increment (annual fuel reduction, power output increase, etc.) in response to design TIAC system cooling capacities. The optimal value of cooling capacity, providing the minimum TIAC system sizes, is associated with the maximum rate of annual effect increment over the overall range of annual effect increment. The rational value of TIAC system design cooling capacity, providing a close to maximum annual effect without TIAC system oversizing, is associated with the second maximum rate of annual effect increment within the range beyond the first maximum rate.

2.2. The Computation Algorithm The cooling potential of GT intake air is estimated by cooling degree hours (CDH) calculated as air temperature depression ∆t = ta ta multiplied by the corresponding duration τ in hours: CDH = − 2 ∆t τ,K h. The summation of current numbers CDH = ∆t τ,K h for a year (month) gives an annual · · · · (monthly) intake air cooling potential: ΣCDH = Σ(∆t τ), K h. · · The real input data on site actual ambient air temperature tamb were taken by using the program “meteomanz” [74]. The use of ambient air cooling potential in terms of CDH provides easy calculation of annual or monthly fuel reduction B or other effect, for instance, annual power energy production or refrigeration energy generation per estimated time period. In the present investigation the climatic characteristics of GT as dependence of specific (or total) fuel consumption be on intake air temperature are used. With this, the CDH numbers are multiplied by the value of decrease ∆be in specific fuel consumption for every 1K drop in intake air temperature, ∆be/∆t, and by turbine power output Pe [61,71]: Energies 2020, 13, 6201 6 of 22

current values of total fuel reduction per an hour •

B = CDH (∆be/∆t) Pe · or specific fuel reduction (for 1 kW turbine power output)

b = CDH (∆be/∆t) (1)

annual total fuel reduction • ΣB = ΣCDH (∆be/∆t) Pe · or annual specific fuel reduction (for 1 kW turbine power output)

Σb = ΣCDH (∆be/∆t) (2)

A total cooling capacity Q0, required for cooling air with flow rate Ga:

Q = ca ξ ∆ta Ga, (3) 0 · · where ∆ta = ta ta —depression of ambient air temperature; t —temperature of cooled air at the − 2 a2 air cooler outlet; ξ—specific heat ratio of the total heat (latent and sensible) rejected from air during cooling to its sensible heat; ca—specific heat of humid (moist) air, kJ/(kg K). · To simplify calculations and to apply their results for any total cooling capacity Q0 and GT power output, it is convenient to carry out the calculations and present their results in relative (specific) value per unit air mass flow rate (Ga = 1 kg/s)—in the form of specific cooling capacity: q0 = P Q0/Ga, kW/(kg/s), or kJ/kg, as well as fuel reduction in specific (relative) values b for 1 kW power output of GT. For simplifying the application of calculation results for GT with various values of decrease ∆be in specific fuel consumption for every 1 ◦C drop in intake air temperature the value of ∆be/∆t = 1.0 g/(kWh K) was assumed. Therefore, if the real value for the concrete GT, for instance, · is ∆be/∆t = 0.35 g/(kWh K), the annual specific fuel reduction Σb (2) should be multiplied by 0.35. · A specific cooling capacity q0, required for cooling air of unit mass flow rate Ga = 1 kg/s:

q = Q /Ga = ca ξ ∆ta, (4) 0 0 · P An improved method of TIAC system rational designed based on the annular fuel reduction B P P curve dependence on cooling capacity Q0 as B = f (Q0) or in specific (relative) values b = f (q0) for 1 kW power output of GT is developed to avoid oversizing. The well-known method of design cooling load calculation based on maximum value of annular P fuel reduction B proceeding from annual CDH number as a primary criterion is used as the first stage of developed methodology. It has been supplemented by addition stages focused to determine a precise value of rational design cooling capacity that enables avoiding TIAC system oversizing. A design specific cooling capacity value q0.opt (corresponding to total cooling capacity value P Q0.opt) that provides a maximum rate b/q0 of annual fuel reduction increment and corresponding P P P annual specific fuel reduction bopt (corresponding to total values B/Q0 and Bopt), and assumed as optimal value, is determined at the second stage of calculation procedure. P The third stage focuses to calculate the maximum rate of annual fuel reduction increment b/q0 but P P P in a remaining range beyond the optimal annual specific fuel reduction bopt:[ b bopt ]/q0, where P P P P P P −P b > bopt or corresponding total values Bopt:[ B Bopt ]/Q0, where B > Bopt. A cooling − P capacity q0.rat, corresponding to the second maximum rate of annual fuel reduction increment b/q0, is considered to be a precise value of rational design specific cooling capacity q0.rat or rational design total cooling capacity value Q0.opt. Energies 2020, 13, 6201 7 of 22

P The fourth stage of proposed methodology is aimed to calculate the annual fuel reduction brat P or Brat close to its maximum value at a precise value of rational design cooling capacity q0.rat or Q0.rat, excluding system oversizing. Details of the calculation procedure have been presented in AppendixA.

Energies3. Results 2020, 13, x FOR PEER REVIEW 7 of 23 The GT operation climatic conditions are characterized by large fluctuations of ambient air parameters:Energies 2020, 13 temperature, x FOR PEER REVIEWtamb, relative ϕamb and absolute damb humidity (Figure1). 7 of 23

Figure 1. Ambient air temperatures tamb, relative 𝑎𝑚𝑏 and absolute damb humidity in 2017, Mykolayiv region (southern Ukraine).

Figure 1. Ambient air temperatures tamb, relative  and absolute damb humidity in 2017, Mykolayiv TheFigure fluctuation 1. Ambient of air ambient temperatures air parameterstamb, relative causesϕamb𝑎𝑚𝑏and large absolute changesdamb ofhumidity specific in cooling 2017, Mykolayiv capacities q0 region (southern Ukraine). requiredregion for (southern cooling ambient Ukraine). air of unit air mass flow rate, Ga = 1 kg/s, at the intake of GT (changes in current specific thermal loads on cooling system) accordingly (Figure 2). TheThe fluctuationfluctuation ofof ambientambient airair parametersparameters causescauses largelargechanges changesof ofspecific specificcooling coolingcapacities capacitiesq q00 The fluctuation of ambient air parameters causes large changes of specific cooling capacities q0 requiredrequired for for cooling cooling ambient ambient air air of of unit unit air air mass mass flow flow rate, rate,G Ga a= = 11 kgkg/s,/s, atat thethe intakeintake ofof GTGT (changes(changes inin required for cooling ambient air of unit air mass flow rate, Ga = 1 kg/s, at the intake of GT (changes in currentcurrent specificspecific thermalthermal loadsloads on on cooling cooling system) system) accordingly accordingly (Figure (Figure2 ).2). current specific thermal loads on cooling system) accordingly (Figure 2). The fluctuation of ambient air parameters causes large changes of specific cooling capacities q0 required for cooling ambient air of unit air mass flow rate, Ga = 1 kg/s, at the intake of GT (changes in current specific thermal loads on cooling system) accordingly (Figure 2).

Figure 2. Changes of specific cooling capacities q0.7, q0.10 and q0.15 required for cooling ambient air with Figure 2. Changes of specific cooling capacities q0.7, q0.10 and q0.15 required for cooling ambient air unit mass flow rate of 1 kg/s from its actual temperature tamb to t2 = 7, 10 and 15 °C. with unit mass flow rate of 1 kg/s from its actual temperature tamb to t2 = 7, 10 and 15 ◦C.

Figure 2. Changes of specific cooling capacities q0.7, q0.10 and q0.15 required for cooling ambient air with SuchSuch large yearly variations ofof specificspecific coolingcooling capacitiescapacitiesq q0.70.7,, qq0.100.10 andand q0.15, ,required required for for cooling cooling unit mass flow rate of 1 kg/s from its actual temperature tamb to t2 = 7, 10 and 15 °C. ambientambient air air from current air temperatures ttaa to target temperaturestemperatures ofof cooledcooled airair tta2a2 = 7,7, 10, 10, andand 15 15 °C◦C revealreveal the the actual actual problem problem to to choose choose a a design design coolin coolingg capacity capacity of of the the chillers chillers and and rational rational chiller chiller Such large yearly variations of specific cooling capacities q0.7, q0.10 and q0.15, required for cooling compoundcompound of of TIAC TIAC system system to to cover cover actual actual cooling cooling duties duties without without its its oversizing oversizing and and provide provide close close to to ambient air from current air temperatures ta to target temperatures of cooled air ta2 = 7, 10, and 15 °C maximummaximum annual annual fuel fuel reduction. reduction. reveal the actual problem to choose a design cooling capacity of the chillers and rational chiller TraditionallyTraditionally a turbine intake air is cooled to the temperature ta2 == 1515 C Cby by a chilled a chilled water water with with a compound of TIAC system to cover actual cooling duties without itsa2 oversizing◦ and provide close to temperaturea temperature of of7 7C fromC from absorption absorption lithium-bromid lithium-bromidee chiller chiller (ACh) (ACh) of of a asimple simple cycle cycle with with a a high high maximum annual fuel◦ reduction. coefficientcoefficient of of performance performance COP from 0.7 to 0.8 (Figure 3 3).). Traditionally a turbine intake air is cooled to the temperature ta2 = 15 C by a chilled water with a temperature of 7 C from absorption lithium-bromide chiller (ACh) of a simple cycle with a high coefficient of performance COP from 0.7 to 0.8 (Figure 3).

Energies 2020, 13, 6201 8 of 22 Energies 2020, 13, x FOR PEER REVIEW 8 of 23

75°C Hot water 90°C Cooling tower

25°C Exhaust gas Ahc

75°C 20°C AC Air 15°C 150°C Exhaust 90°C Boiler

30°C GT Electrical 12°C Generator Cooled water 7°C

12°C 7°C Electricity

FigureFigure 3. 3. AA scheme scheme of of gas gas turbine turbine intake intake air air cooling system with with absorption lithium-bromide lithium-bromide chiller chiller (ACh)(ACh) using using the the heat heat of of exhaust exhaust gas: AC—air AC—air cooler. cooler.

AsAs mentioned above, thethe lowestlowest temperaturetemperature of of air air cooled cooled in in ACh ACh of of a simplea simple cycle cycle is aboutis about 15 15◦C °Cand and limited limited by by the the temperature temperature of chilledof chilled water water of 7of◦ C7 °C at theat the inlet inlet of airof air cooler. cooler. ToTo provide turbine intake air cooling to lower temperature of about 10 °C◦C with corresponding incrementincrement of of fuel fuel saving saving the the refrigerant refrigerant with with lower lower boiling boiling point point can can be be used used as as a a coolant coolant in in the the inlet inlet airair cooler cooler and and the the ECh ECh applied applied as as the the most most simple simple in in design design and and cheap, cheap, but but not not so so high high efficient efficient as as ACh: its its COP COP is is 0.2 0.2 to to 0.3 0.3 versus versus COP COP of of 0.7 0.7 for for ACh. ACh. SoSo as as the the COP COP of of ECh ECh falls falls with with lowering lowering the the boiling boiling temperature temperature of of refrigerant, refrigerant, in in order order to to keep keep thethe COP COP at at the the level level of of about 0.3 0.3 the refrigerant boiling temperature temperature is is to to be about t00 fromfrom 4 4 to to 5 5 °C◦C andand a a thermal thermal load load on on ECh ECh should should be be restricted restricted by by its its range, range, referred referred to to subcooling subcooling the the air, air, previously previously precooledprecooled in high efficient efficient ACh. IssuingIssuing from from this this approach, approach, a a novel novel two-stage two-stage TIAC TIAC system system has has been been proposed proposed (Figure (Figure 44).). Apparently it is reasonable to cool intake air from current ambient air temperature tamb toto tat2a 2= =1515 C ◦inC AChin ACh with with high high coefficient coefficient of perf of performanceormance COP COP from from 0.7 to 0.7 0.8, to 0.8,but further but further deep deep cooling cooling air from air from ta2 = 15ta2 °C= 15 down◦C down to ta2to = t10a2 =°C10 is◦ Cpossible is possible by boiling by boiling refriger refrigerantant from from ECh, ECh, i.e., i.e., in in combined combined two-stage two-stage absorption-ejectorabsorption-ejector chiller (AECh) (Figure 44).). The use of ECh as a second stage of TIAC with previous ambientambient airair precoolingprecooling is is caused caused by by its lowits low coe fficoefficientcient of performance of performance COP fromCOP 0.2from to 0.30.2 that to 0.3 requires that requiresenlarged enlarged values of values heat generated of heat generated in exhaust in boilerexhaust in boiler the form in the of hotform water of hot or water steam. or steam. Further enhancing the efficiency of TIAC systems and the effect, as turbine fuel saving for example, gained due to their application, especially in the case of their running in temperate climatic conditions, is still depended on lowering the intake air temperature, limited by temperature of coolant in the air cooler. In the case of ECh, as low-temperature stage of two-stage AECH (SAECh), a decrease in refrigerant boiling temperature to t0 of 2 to 3 ◦C is accompanied by dropping its COP to 0.2 and lower. To compensate for a negative effect of decreasing a refrigerant boiling temperature t0 in ejector thermodynamic cycle, the refrigerant condensing temperature tc should be decreased too. This might be realized through cooling the condenser of ECh by chilled water from ACh with temperature of about 7 ◦C or by returned chilled water leaving the intake air cooler (high-temperature stage ACHT of hybrid intake air cooler) with temperature of about 10 to 12 ◦C. The last variant is more efficient due to the use of chilled water of enlarged flow rate in intake air cooler that is accompanied by decreasing its temperature arise and leads to increasing a heat flux in air cooler and reducing its dimensions as a result.

Energies 2020, 13, 6201 9 of 22

Energies 2020, 13, x FOR PEER REVIEW 9 of 23

75°C 90°C Hot water

Exhaust gas Ahc Cooling tower

75°C

AC AC 90°C HТ LТ 150°C Air 10°C Exhaust boiler Electrical Cooled water 7°C Generator 30°C Ref rig. GT Generator 12°C 85°C Ejector 40°C Ech Condenser Exp. valve 3°C0 3°C 8°C 7°C

12°C Electricity Figure 4.4. AA schemescheme of of two-stage two-stage gas gas turbine turbine intake intake air coolingair cooling system system in stage in absorption-ejectorstage absorption-ejector chiller

(SAECh):chiller (SAECh): ACHT—high-temperature ACHT—high-temperature stage ofstage air cooler;of air cooler; ACLT—low-temperature ACLT—low-temperature stage ofstage air cooler;of air Exp.Valve—expansioncooler; Exp.Valve—expansion valve. valve.

TheFurther reserves enhancing for the the use efficiency of cooling of capacityTIAC system of AChs and for the enhancing effect, as the turbine efficiency fuel ofsaving ACh arefor proceededexample, gained from due the variationto their applic of currentation, especially thermal loads in the on case TIAC of their system. running At in lowered temperate ambient climatic air temperatures,conditions, is andstill thermaldepended load on on lowering TIAC system the intake accordingly, air temperature, the excessive limited cooling by capacitytemperature of ACh of cancoolant be used in the for air cooling cooler. the In condenserthe case of ofECh, ECh as to low-temperature provide deeper airstage cooling of two-stage to 7 ◦C andAECH lower (SAECh), due to lowereda decrease refrigerant in refrigerant boiling boiling temperature temperaturet0 of to 2 tot0 3of◦ 2C to as 3 compared °C is accompanied with two-stage by dropping cooling its air COP to the to temperature0.2 and lower. of To 10 compensate◦C (at t0 of 4 for to 5a◦ negativeC) in a stage effect absorption-ejector of decreasing a refrigerant chiller (SAECh). boiling temperature t0 in ejectorIn this thermodynamic case, the ECh operates cycle, the as refrigerant the low cascade condensing of cascade temperature absorption-ejector tc should chiller be decreased (CAECh) too. to provideThis might deeper be realized cooling thethrough intake cooling air (to 7 the◦C andcondenser lower) previouslyof ECh by precooledchilled water to the from temperature ACh with of 15Energiestemperature◦C in 2020 the, 13 first of, x FORabout high-temperature PEER 7 °C REVIEW or by returned stage ofchilled air cooler water by leaving ACh (Figure the intake5). air cooler (high-temperature10 of 23 stage ACHT of hybrid intake air cooler) with temperature of about 10 to 12 °C. The last variant is more 75°C efficient due to the use of chilled water of enlarged flow rate in intake air cooler thatHot is water accompanied by decreasing its temperature arise and leads to increasing a heat flux in air cooler and reducing90°C its dimensions as a result. Cooling tower The reserves for the use of cooling capacity of ACh for enhancing the efficiency of ACh are proceeded from the variation of currentExhaust gas thermal loads on TIAC system. At lowered ambient air temperatures, and thermal load on TIAC system accordingly, the excessive cooling capacity of ACh can be used for cooling the condenser of ECh to provide deeper air cooling to 7 °C and lower due to 75°C lowered refrigerant boiling temperature t0 of 2 to 3 °C as compared with two-stage cooling air to the ACHTACLT 90°C temperatureAir of 10 °C (at t0 of 4 to 5 °C) in a stage absorption-ejector chiller (SAECh). 10°C 150°C Exhaust In this case, the ECh operates as the Boilerlow cascadElectricale of cascade absorption-ejector chiller (CAECh) Cooled water 7°C to provide deeper cooling the intake air (to 7 °C andGenerator lower) previously precooled to the temperature 30°C GT 12°C of 15 °C in the first high-temperature stage of air cooler by ACh85°C (FigureEjector 5). 40°C Refrig. EhC Condenser Generator 20°C 3°C 8°C 7°C 12°C Electricity

Figure 5. A schemescheme of of two-stage two-stage gas gas turbine turbine intake intake air coolingair cooling system system in stage in stage cascade cascade absorption-ejector absorption- ejectorchiller (SCAECh).chiller (SCAECh).

Such functioning of ECh as the low-temperature stage in SAECh to provide intake air cooling to 10 °C at increased ambient air temperatures and as low cascade of CAECh to provide deeper intake air cooling to 7 °C and lower at decreased ambient air temperature enables matching daily and seasonal fluctuations in thermal loads and efficient operation of TIAC systems in actual site climatic conditions. In this case, we deal with universal combined stage-cascade absorption-ejector chiller (SCAECh), functioning as SAECh at increased ambient air temperatures and as SCAECh at its decreased temperatures or even as CAECh at ambient air temperatures close to 15 °C and lower. Such SCAECh provides a maximum effect, gained due to deep TIAC and matching daily and seasonal fluctuations in thermal loads, especially in temperate climatic conditions. To determine a rational design thermal load on TIAC system that provides close to maximum annual values of fuel reduction the characteristic curves of annual values of CDH versus specific cooling capacity q0 (per unit air mass flow rate Ga = 1 kg/s) and total fuel reduction ∑B versus overall cooling capacity Q0 (per total air mass flow rate Ga) to target temperatures of cooled air ta2: 7 and 10 °C—in AECh; 15 °C—in ACh for climatic conditions in Mykolayiv region (southen Ukraine) are calculated as the first, traditional, stage of TIAC system designing (Figure 6).

ΣCDH, °C·h ΣB, t 70,00070000 700 ∑CDH7 ∑B7 60,00060000 600 50,00050000 ∑B10 ∑CDH10 500 40,00040000 400

30,00030000 300 ∑B15 ∑CDH15 20,00020000 200 10,00010000 100

00 0 0 5 10 15 20 25 30 35q 40, kW/(kg/s) 45 50 0 500 1000 1500 2000Q 25000, kW 0 (a) (b)

Figure 6. The annual values of CDH versus specific cooling capacity q0 (a) and total fuel reduction ∑B

versus overall cooling capacity Q0 (b) for ta2: 7 and 10 °C—in AECh; 15 °C—in ACh.

Energies 2020, 13, x FOR PEER REVIEW 10 of 23

75°C Hot water 90°C Cooling tower

Exhaust gas

75°C AC AC Air HT LT 90°C 10°C 150°C Exhaust Boiler Electrical Cooled water Generator 7°C 30°C GT 12°C 85°C Ejector 40°C Refrig. EhC Condenser Generator 20°C 3°C 8°C 7°C 12°C Electricity Figure 5. A scheme of two-stage gas turbine intake air cooling system in stage cascade absorption- Energies 2020, 13, 6201 10 of 22 ejector chiller (SCAECh).

SuchSuch functioningfunctioning ofof EChECh asas thethe low-temperaturelow-temperature stagestage inin SAEChSAECh toto provideprovide intakeintake airair coolingcooling toto 1010◦ °CC atat increasedincreased ambient ambient air air temperatures temperatures and and as as low low cascade cascade of CAEChof CAECh to provideto provide deeper deeper intake intake air coolingair cooling to 7 ◦toC and7 °C lowerand lower at decreased at decreased ambient ambient air temperature air temperature enables matchingenables matching daily and daily seasonal and fluctuationsseasonal fluctuations in thermal in loads thermal and loads efficient and operation efficient ofoperation TIAC systems of TIAC in actualsystems site in climatic actual site conditions. climatic conditions.In this case, we deal with universal combined stage-cascade absorption-ejector chiller (SCAECh), functioningIn this case, as SAECh we deal at with increased universal ambient combined air stage-cascade temperatures absorption-ejector and as SCAECh chiller at its (SCAECh), decreased temperaturesfunctioning as or evenSAECh as CAEChat increased at ambient ambient air temperaturesair temperatures close and to 15 as◦C SCAECh and lower. at Suchits decreased SCAECh providestemperatures a maximum or even e asffect, CAECh gained at dueambient to deep air TIACtemperatures and matching close to daily 15 °C and and seasonal lower. fluctuationsSuch SCAECh in thermalprovides loads, a maximum especially effect, in temperate gained due climatic to deep conditions. TIAC and matching daily and seasonal fluctuations in thermalTo determine loads, especially a rational in design temperate thermal climatic load conditions. on TIAC system that provides close to maximum annualTo valuesdetermine of fuel a rational reduction design the characteristicthermal load on curves TIAC of system annual that values provides of CDH close versus to maximum specific P coolingannual capacityvalues ofq 0fuel(per reduction unit air mass the flowcharacteristic rate Ga = 1curves kg/s) andof annual total fuel values reduction of CDHB versus versus specific overall coolingcooling capacitycapacity Qq00 (per(per unit total air air mass mass flow flow rate rate Ga G= a1) kg/s) to target and temperaturestotal fuel reduction of cooled ∑B versus air ta2: overall 7 and 10cooling◦C—in capacity AECh; Q 150 ◦(perC—in total ACh air formass climatic flow rate conditions Ga) to target in Mykolayiv temperatures region of (southencooled air Ukraine) ta2: 7 and are 10 calculated°C—in AECh; as the 15 first, °C—in traditional, ACh for stage climatic of TIAC conditions system designingin Mykolayiv (Figure region6). (southen Ukraine) are calculated as the first, traditional, stage of TIAC system designing (Figure 6).

ΣCDH, °C·h ΣB, t 70,00070000 700 ∑CDH7 ∑B7 60,00060000 600 50,00050000 ∑B10 ∑CDH10 500 40,00040000 400

30,00030000 300 ∑B15 ∑CDH15 20,00020000 200 10,00010000 100

00 0 0 5 10 15 20 25 30 35q 40, kW/(kg/s) 45 50 0 500 1000 1500 2000Q 25000, kW 0 (a) (b) P FigureFigure 6.6. TheThe annualannual valuesvalues ofof CDHCDH versusversus specificspecific coolingcooling capacity capacityq q00( (aa)) andand totaltotal fuelfuel reductionreduction ∑BB versusversus overalloverall coolingcooling capacitycapacityQ Q00 ((bb)) forfor ttaa22:: 7 7 and and 10 10 °C—in◦C—in AECh; AECh; 15 15 °C—in◦C—in ACh. ACh.

The calculations are carried out for GT of rated power output NeISO = 10 MW and with account to assumption that a reduction of air temperature ∆ta by 1K leads to a decrease in specific fuel P P consumption ∆be by 1.0 g/(kWh), as it was assumed for easy calculation of fuel reduction B and b for another value of ∆be/∆ta. As it was mentioned above, in order to simplify calculations and to apply their results for any total cooling capacity Q0 and GT power output, it is convenient to carry out the calculations and present their results in relative (specific) value per unit air mass flow rate (Ga = 1 kg/s)—in the form of specific cooling capacity: q0 = Q0/Ga, kW/(kg/s), or kJ/kg, as well as annual fuel reduction in specific (relative) P values b for 1 kW power output of GT. P The annual values of GT specific fuel reduction ∆b (for 1 kW of GT power) due to intake air cooling from ambient air temperatures tamb to ta2 = 7, 10 and 15 ◦C by chillers with various specific cooling capacities q0 (per unit air mass flow rate Ga = 1 kg/s) for climatic conditions in Mykolayiv region (southen Ukraine) are presented in Figure7. Figures6 and7 show that intake air cooling in AECh to ta2 = 10 ◦C provides more than 50% greater annual reduction in fuel consumption compared to cooling air to ta2 = 15 ◦C in ACh. As Figure7 shows, a specific cooling capacity q0.10 is 33 to 35 kW/(kg/s), or kJ/kg, provide cooling ambient air from P the current temperatures ta to ta2 = 10 ◦C with annual specific fuel reduction b10 of 47 to 48 kg/kW (Figure7) with corresponding overall cooling capacity Q0.10 of 1300 to 1400 kW providing a total fuel P reduction B10 of 470 to 480 t (Figure6) that is close to their maximum value at noticeable high rate of their annual increment. While beyond the values Q0.10 of 1300 to 1400 kW or q0.10 of 33 to 35 kW/(kg/s) Energies 2020, 13, x FOR PEER REVIEW 11 of 23

The calculations are carried out for GT of rated power output NeISO = 10 MW and with account to assumption that a reduction of air temperature Δta by 1K leads to a decrease in specific fuel consumption Δbe by 1.0 g/(kWh), as it was assumed for easy calculation of fuel reduction ∑B and ∑b for another value of Δbe/Δta. As it was mentioned above, in order to simplify calculations and to apply their results for any total cooling capacity Q0 and GT power output, it is convenient to carry out the calculations and present their results in relative (specific) value per unit air mass flow rate (Ga = 1 kg/s)—in the form Energiesof specific2020 ,cooling13, 6201 capacity: q0 = Q0/Ga, kW/(kg/s), or kJ/kg, as well as annual fuel reduction in specific11 of 22 (relative) values ∑b for 1 kW power output of GT. The annual values of GT specific fuelP reductionP ∑Δb (for 1 kW of GT power) due to intake air a rate of annual fuel reduction increment B or b is negligible, the range of corresponding cooling cooling from ambient air temperatures tamb to ta2 = 7, 10 and 15 °C by chillers with various specific capacities Q0 or q0 needed is wide that approves a considerable chiller oversizing. Thus, the precise cooling capacities q0 (per unit air mass flow rate Ga = 1 kg/s) for climatic conditions in Mykolayiv values of design cooling capacities Q or q are to be determined for appropriately sized TIAC system. region (southen Ukraine) are presented0 in0 Figure 7.

Σb, g/kW 70,00070000 Σb7 60,00060000 Σb 50,00050000 10 40,00040000 30,00030000 Σb15 20,00020000 10,000 10000 0 0 q , kW/(kg/s) 0 5 10 15 20 25 30 35 40 450 50 55

Figure 7. The annual values of specific fuel reduction P∑Δbe against specific cooling capacity q0 at Figure 7. The annual values of specific fuel reduction ∆be against specific cooling capacity q0 at different temperatures of cooled air ta2: 7 and 10 °C—in AECh; 15 °C—in ACh. different temperatures of cooled air ta2: 7 and 10 ◦C—in AECh; 15 ◦C—in ACh.

a2 Furthermore,Figures 6 and with 7 show arising that a targetintake temperatureair cooling in of cooledAECh to air tta 2=to 10 15 °C◦C provides it is more more problematic than 50% to a2 selectgreater precise annual value reduction of rational in fuel design consumption cooling compared capacity Q to0.15rat coolingor q air0.15rat to tso = as 15 behavior °C in ACh. of As the Figure curve P P 0.10 7 shows,B = f (Q a0 )specific or b =coolingf (q0) becomes capacity moderateq is 33 to (Figures 35 kW/(kg/s),6 and7). or Therefore kJ/kg, provide the original cooling methodambient of air determiningfrom the current a design temperatures cooling capacity ta to ta2has = 10 been °C with modified annual by addingspecific thefuel stages reduction aimed ∑ forb10 calculationof 47 to 48 0.10 ofkg/kW optimal (Figure design 7) with cooling corresponding capacity Q0.opt overallor q 0.optcoolingcorresponding capacity Q to of a maximum1300 to 1400 rate kW of providing annual fuel a P 10 P reductiontotal fuel reduction increment ∑BB/Q of0 or470 tob/ q4800 as tintermediate (Figure 6) that stage is close (Figure to 8their) for maximum determining value theprecise at noticeable value · 0.10 0.10 ofhigh rational rate of design their annual specific increment. cooling capacity While qbeyond0.rat (Figures the values9–11). Q of 1300 to 1400 kW or q of 33 to 35 kW/(kg/s)The points a Orate15 and of Oannual10 on the fuel TIAC reduction cumulative increment characteristics ∑B or of∑b dependence is negligible, of annual the range specific of P P 0 0 fuelcorresponding reduction coolingb15 and capacitiesb10 on the Q specific or q cooling needed capacities is wide qthat0.15 and approvesq0.10 are a determined considerable according chiller 0 0 tooversizing. optimal values Thus, ofthe design precise specific values cooling of design capacities coolingq0.15opt capacitiesand q 0.10optQ or, q corresponding are to be determined to maximum for P P valuesappropriately of ratios sizedB/Q TIAC0 or system.b/q0, i.e., maximum rate of annual fuel reduction increment. · a2 P AsFurthermore, Figure8 shows, with arising a maximum a target rate temperature of annual of specific cooled fuel air t reduction to 15 °C it increment is more problematicb/q0.15 due to select precise value of rational design cooling capacity Q0.15rat or q0.15rat so as behavior of the curve ∑B to cooling air to ta2 = 15 ◦C takes place at the optimal design specific cooling capacity q0.15opt = = f(Q0) or ∑b = f(q0) becomes moderate (Figures 6 and 7). ThereforeP the original method of determining 16 kW/(kg/s) and provides the annual specific fuel reduction b15opt = 22.5 kg/kW considerably less a design cooling capacityP has been modified by adding the stages aimed for calculation of optimal than its maximum value b15max = 27 kg/kW. With cooling air targeting ta2 = 10 ◦C the optimal specific design cooling capacity Q0.opt or q0.opt correspondingP to a maximum rate of annualP fuel reduction cooling capacity q0.10opt = 27 kW/(kg/s) provides b10opt = 42 kg/kW less versus b10max = 49 kg/kW. Energies 2020, 13, x FOR PEER REVIEW 12 of 23 increment ∑B/Q0·or ∑b/q0 as intermediate stage (Figure 8) for determining the precise value of rational design specific cooling capacity q0.rat (Figures 9–11). 1, 2st Σb, g/kW 1, 2st Σb, g/kW Σb/q0,g/kW/(kg/s) Σb/q0,g/kW/(kg/s) 1400 30,00030000 1800 60,00060000 t = 15°C ∑b t =10°C 2 15max 1600 2 1200 25,000 Σb/q0.10 ∑b10max 50,000 Σb/q0.15 ∑b 25000 50000 15opt 1400 ∑b 1000 O O10 15opt 15 20,00020000 1200 40,00040000 ∑b 800 1000 10 15,00015000 30,00030000 600 800 10,00010000 600 20,00020000 400 400 5,0005000 10,00010000 200 200 0 00 0 00 q15opt q10opt q , kW/(kg/s) 0 5 10 15 20 25 30 35 40 45 500 55 0 5 10 15 20 25 30 35 40 45q 500, kW/(kg/s) 55 (a) (b) P P Figure 8. Annual specificspecific fuelfuel reduction reduction ∑bband and its its relative relative increment increment b∑/qb0/qdue0 due to to cooling cooling ambient ambient air P toairt toa2 =ta215 = 15◦C( °Ca )(a and) and 10 10◦C( °Cb )(b against) against specific specific cooling cooling capacity capacityq0 :q0q:0.opt q0.optand and ∑bboptopt—optimal values.

The points O15 and O10 on the TIAC cumulative characteristics of dependence of annual specific fuel reduction ∑b15 and ∑b10 on the specific cooling capacities q0.15 and q0.10 are determined according to optimal values of design specific cooling capacities q0.15opt and q0.10opt, corresponding to maximum values of ratios ∑B/Q0·or ∑b/q0, i.e., maximum rate of annual fuel reduction increment. As Figure 8 shows, a maximum rate of annual specific fuel reduction increment ∑b/q0.15 due to cooling air to ta2 = 15 °C takes place at the optimal design specific cooling capacity q0.15opt = 16 kW/(kg/s) and provides the annual specific fuel reduction ∑b15opt = 22.5 kg/kW considerably less than its maximum value ∑b15max = 27 kg/kW. With cooling air targeting ta2 = 10 °C the optimal specific cooling capacity q0.10opt = 27 kW/(kg/s) provides ∑b10opt = 42 kg/kW less versus ∑b10max = 49 kg/kW. To achieve close to maximum annual specific fuel reduction ∑b15opt at more precise value of rational cooling capacity q0.rat without oversizing the chiller, it is proposed to determine the maximum rate of annual specific fuel reduction increment as (∑b − ∑bopt)/q0.15 in the range of ∑b15 beyond its value ∑b15opt = 22.5 kg/kW corresponding to optimal design capacity q0.15opt = 16 kW/(kg/s) for cooling air to ta2 = 15 °C as well as beyond ∑b10opt = 42 kg/kW corresponding to q0.10opt = 27 kW/(kg/s) for ta2 = 10 °C (Figures 9 and 10).

Σb/q0,10(Σb-Σbopt)/q0, g/kW/(kg/s) 2, 3st Σb/q0,10(Σb-Σbopt)/q0, g/kW/(kg/s) 2, 3st 2000 1800 t2=15°C R t2=10°C 15 10(Σb-Σbopt)/q0.15 1800 1600 O10 1600 R 10(Σb-Σbopt)/q0.10 1400 Σb/q 10 1400 0.10 Σb/q O 1200 1200 0.15 15 1000 1000 800 800 600 600 400 400 200 200 0 0 q q15rat q10opt q10rat 15opt q , kW/(kg/s) q , kW/(kg/s) 0 5 10 15 20 25 30 35 40 450 50 55 0 5 10 15 20 25 30 35 40 450 50 55 (a) (b)

Figure 9. Relative increments of annual specific fuel reduction ∑b/q0 and (∑b − ∑bopt)/q0 for ta2 = 15 °C

(a) and 10 °C (b) against specific cooling capacity q0.

Energies 2020, 13, x FOR PEER REVIEW 12 of 23

1, 2st Σb, g/kW 1, 2st Σb, g/kW Σb/q0,g/kW/(kg/s) Σb/q0,g/kW/(kg/s) 1400 30,00030000 1800 60,00060000 t = 15°C ∑b t =10°C 2 15max 1600 2 1200 25,000 Σb/q0.10 ∑b10max 50,000 Σb/q0.15 ∑b 25000 50000 15opt 1400 ∑b 1000 O O10 15opt 15 20,00020000 1200 40,00040000 ∑b 800 1000 10 15,00015000 30,00030000 600 800 10,00010000 600 20,00020000 400 400 5,0005000 10,00010000 200 200 0 00 0 00 q15opt q10opt q , kW/(kg/s) 0 5 10 15 20 25 30 35 40 45 500 55 0 5 10 15 20 25 30 35 40 45q 500, kW/(kg/s) 55 (a) (b)

Figure 8. Annual specific fuel reduction ∑b and its relative increment ∑b/q0 due to cooling ambient

air to ta2 = 15 °C (a) and 10 °C (b) against specific cooling capacity q0: q0.opt and ∑bopt—optimal values.

The points O15 and O10 on the TIAC cumulative characteristics of dependence of annual specific fuel reduction ∑b15 and ∑b10 on the specific cooling capacities q0.15 and q0.10 are determined according to optimal values of design specific cooling capacities q0.15opt and q0.10opt, corresponding to maximum values of ratios ∑B/Q0·or ∑b/q0, i.e., maximum rate of annual fuel reduction increment. As Figure 8 shows, a maximum rate of annual specific fuel reduction increment ∑b/q0.15 due to cooling air to ta2 = 15 °C takes place at the optimal design specific cooling capacity q0.15opt = 16 kW/(kg/s) and provides the annual specific fuel reduction ∑b15opt = 22.5 kg/kW considerably less than its maximum value ∑b15max = 27 kg/kW. With cooling air targeting ta2 = 10 °C the optimal specific cooling capacity q0.10opt = 27 kW/(kg/s) provides ∑b10opt = 42 kg/kW less versus ∑b10max = 49 kg/kW. To achieve close to maximum annual specific fuel reduction ∑b15opt at more precise value of rational cooling capacity q0.rat without oversizing the chiller, it is proposed to determine the maximum rate of annual specific fuel reduction increment as (∑b − ∑bopt)/q0.15 in the range of ∑b15 beyond its value ∑b15opt = 22.5 kg/kW corresponding to optimal design capacity q0.15opt = 16 kW/(kg/s) for cooling Energies 2020, 13, 6201 12 of 22 air to ta2 = 15 °C as well as beyond ∑b10opt = 42 kg/kW corresponding to q0.10opt = 27 kW/(kg/s) for ta2 = 10 °C (Figures 9 and 10).

Σb/q0,10(Σb-Σbopt)/q0, g/kW/(kg/s) 2, 3st Σb/q0,10(Σb-Σbopt)/q0, g/kW/(kg/s) 2, 3st 2000 1800 t2=15°C R t2=10°C 15 10(Σb-Σbopt)/q0.15 1800 1600 O10 1600 R 10(Σb-Σbopt)/q0.10 1400 Σb/q 10 1400 0.10 Σb/q O 1200 1200 0.15 15 1000 1000 800 800 600 600 400 400 200 200 0 0 q q15rat q10opt q10rat 15opt q , kW/(kg/s) q , kW/(kg/s) 0 5 10 15 20 25 30 35 40 450 50 55 0 5 10 15 20 25 30 35 40 450 50 55 (a) (b) P P P Figure 9. RelativeRelative increments increments of of annu annualal specific specific fuel fuel reduction reduction ∑b/qb0/ qandand (∑b ( − b∑bopt)/bqopt0 for)/q ta2for = 15ta °C= 0 − 0 2 Energies15(a 2020) ◦andC(, 13a 10), andx °C FOR ( 10b PEER)◦ againstC(b REVIEW) against specific specific cooling cooling capacity capacity q0. q0. 13 of 23

4 st Σb, g/kW 4 st Σb, g/kW 10(Σb-Σbopt)/q0, g/kW/(kg/s) 10(Σb-Σbopt)/q0, g/kW/(kg/s) 2000 30000 1400 60,00060000 t2=15°C ∑b 30,000 R15 15max t2=10°C 1800 10(Σb-Σbopt)/q0.10 10(Σb-Σbopt)/q0.15 ∑b 25000 1200 R ∑b10max 50,000 1600 O 15rat 25,000 10 50000 15 O10 ∑b 1400 1000 10rat ∑b15opt 20000 40,00040000 Σb15 20,000 Σb ∑b10max 1200 800 10 1000 15,00015000 30,00030000 800 600 Δq0.15 Δq0.10 20,00020000 600 10,00010000 400 400 10,000 5,0005000 200 10000 200 0 0 0 0 0 q15opt q15rat q15 max 0 q10opt q10rat q10max q , kW/(kg/s) 0 5 10 15 20 25 30 35 40 45q 500, kW/(kg/s) 55 0 5 10 15 20 25 30 35 40 45 500 55 (a) (b) P FigureFigure 10. 10. AnnualAnnual specific specific fuel fuel reduction reduction ∑b bandand relative relative increment increment of of a annualnnual specific specific fuel fuel reduction reduction P P P P Energies(∑( b2020 b− ∑, b13optb,opt )/x qFOR)0/ qfor forPEER ta2 t=a 15REVIEW= °C15 (aC() anda) and 10 °C 10 (bC():·−b():∑b −( ∑bbopt)/qb0opt. )/q . 14 of 23 − 0 2 ◦ ◦ ·− − 0

Here: q0.opt—optimal, q0.rat—rational and q0.max2st—maximum values;Σ pointsb, g/kW O15 and O10 correspond Σb/q0,g/kW/(kg/s) to q0.15opt and q0.10opt; R15 and R10—for q0.15rat and q0. 10rat; Δq0.15 = q0.15max − q0.15rat and Δq0.10 = q0.10max − q0.10rat − 1800 Σb/q ∑b 70,00070000 t2=7,10,15°C 0.7 7 reduction of design specific1600 cooling capacity q0. Σb/q0.10 O7 ∑b7opt 60,00060000 As Figures 9 and 101400 show, a maximum rate of annual ∑specificb fuel reduction increment (∑b − Σb/q0.15 10 50,00050000 opt 0.15 1200 a2 ∑b10opt 0.15rat ∑b )/q when cooling ambient air to t = 15 O°C10 takes place at specific cooling capacity q = 24 kW/(kg/s) and provides1000 annual specific fuel reduction ∑b15rat very close 40,000to40000 its maximum value ∑b15max

= 27 kg/kW, but achieved800 at a rational cooling capacity∑b15 q0.15rat = 24 kW/(kg/s)30,00030000 that is less than its ∑b15opt 600 O15 oversized maximum value q0.15max = 31 kW/(kg/s) by Δq0.15 = q0.15max − q0.15rat20,00020000 of about 7 kW/(kg/s), i.e., 400 more than 20%. When cooling ambient air to ta2 = 10 °C a maximum value of (∑b − ∑bopt)/q0.10 takes 200 10,00010000 place at the rational cooling capacity q0.10rat = 37 kW/(kg/s) and provides annual specific fuel reduction 0 q q 0 0 ∑b10rat = 47 kg/kW that is close to its maximum15opt value10optq 7opt∑b10max = 48 kg/kW and achieved at a rational 0 5 10 15 20 25 30 35 40 45q 500, kW/(kg/s) 55 cooling capacity q0.10rat = 37 kW/(kg/s) less than its oversized maximum value q0.10max = 44 kW/(kg/s) by

Δq0.10 Figure= q0.10max 11. − qRelative0.10rat of aboutincrement 9 kW/(kg/s), of annual i.e., specific approximately fuel reductionP 20% less.∑b/q 0The and points annual R 15specific and R 10fuel on the Figure 11. Relative increment of annual specific fuel reduction b/q0 and annual specific fuel reduction TIACreduction Pcumulative ∑b duecharacteristics to cooling air of to depend ta2 = 7, 10ence and of15 annual°C against specific a cooling fuel capacity reduction needed ∑b15 q 0= (2 f( qstage0) and of ∑b10 b due to cooling air to ta2 = 7, 10 and 15 ◦C against a cooling capacity needed q0 (2 stage of = f(q0)methodology).methodology). on the specific cooling capacities q0.15 and q0.10 are determined according to rational values of design specific cooling capacities q0.15rat and q0.10rat, calculated at the stage 3 of design methodology. As P it is Asseen,To Figures achieve the values 11 close and to of12 maximum rationalshow, a maximumannual annual specific specific rate of fuelannual reductionreduction fuel reduction ∑bb15opt15rat andincrementat more ∑b10rat precise ,( ∑determinedb − value∑bopt)/ ofq0 proceedinginrational the ranges cooling from beyond capacitythe points theirq0.rat Roptimal15without and R values10, oversizing are very∑bopt close(Figure the chiller, to their11) itis maximumis characterized proposed values to determineby ∑ maximumb15max and the maximum∑valuesb10max. of P P P ratiosrateGraphs of (∑ annualb − ∑ inbopt specificFigures)/q0 which fuel9 and correspond reduction 10 allow increment usersto rational to match as design ( bcurrent specificbopt cooling)/ coolingq0.15 induties thecapacities range and provide ofq0.rat bwhen15 minimumbeyond cooling its P − sizingambientvalue systemb air15opt to = atthe22.5 maximum temperatures kg/kW corresponding rate tofa2 =annular 7, 10to and optimalcoolin 15 °C:g designeffect q0.7rat = capacityincrement 41 kW/(kg/s),q0.15opt due q=to0.1016 operationrat kW = 35/(kg kW/(kg/s)/s) at for optimal cooling and P coolingqair0.15rat to =ta 242capacity= kW/(kg/s)15 ◦C q as0.10 wellopt (stage (Figure as beyond3, Figure 9) or bto12).10opt peak These= 42cooling kgvalues/kW needs correspondingof rational without design tooversizingq0.10opt specific= due27 cooling kW to/ (kgrunning capacities/s) for atta2 rationalq=0.rat10 determine◦C cooling (Figures thecapacity9 and position 10 q).0.rat of(Figure the points 10). R7, R10 and R15 on the TIAC cumulative characteristics of dependence of annual specific fuel reduction ∑b7 = f(q0), ∑b10 = f(q0) and ∑b15 = f(q0) and corresponding 4.rational Discussion annual specific fuel reduction ∑b7rat, ∑b10rat and ∑b15rat, calculated at the stage 4 of design methodology (Figure 13). To generalize the results of analysing the efficiency of engine intake air cooling in TIAC systems, using chillers of different types, the optimal and rational design thermal loads for cooling intake air to target temperatures 7Σ bto/q 015,10(Σ °Cb -Σinb oactualpt)/q0, g/ sitekW/(kg/s) climatic conditions3st were calculated step-by-step year 2000 round on the basis of the proposedt2=7,10,15°C novel methodologyR15 (Figures 11–14). 1800 O7 As Figure 11 shows, a maximum rate of annual specificR fuel reduction ∑b/q0 due to cooling 1600 7 ambient air to the temperatures ta2 = 7, 10 and 15 °C takesR10 place at the optimal design specific cooling 1400 O10 O15 capacities (points O7, O120010, O15): q0.7Σoptb/ q=0.7 33 kW/(kg/s), q0.10opt = 27 kW/(kg/s) and q0.15opt = 16 kW/(kg/s), Σb/q which provide annual1000 specific fuel0.10 reduction ∑b7opt = 56 kg/kW, ∑b10opt = 42 kg/kW and ∑b15opt = 22 Σb/q0.15 kg/kW considerably less800 than their maximum values. These values of ∑bopt are used to determine the range of annual specific600 fuel reduction (∑b − ∑bopt) beyond their optimal values ∑bopt to calculate the 400 ratios (∑b − ∑bopt)/q0 that characterize the rate of annual specific fuel reduction beyond the optimal 200 q values ∑bopt and their maximum values correspond to7opt rational design specific cooling capacities: q0.rat. 0 q 15opt q15rat q10opt q10rat q7rat 0 5 10 15 20 25 30 35 40 45q , kW/(kg/s) 50 55 0

Figure 12. Relative increments of annual specific fuel reduction ∑b/q0 and (∑b − ∑bopt)/q0·due to cooling air ta2 = 7, 10 and 15 °C against specific cooling capacities needed q0 (3 stage): −∑b/q0; −(∑b − ∑bopt)/q0.

As it is seen, the values of rational annual specific fuel reduction ∑b7rat, ∑b10rat and ∑b15rat, determined proceeding from the points R7, R10 and R15, are very close to their maximum values. As Figure 13 shows, a maximum rate of annual specific fuel reduction increment (∑b − ∑bopt)/q0.15 when cooling ambient air to ta2 = 15 °C takes place at specific cooling capacity q0.15rat = 24 kW/(kg/s) and provides annual specific fuel reduction ∑b15rat very close to its maximum value ∑b15max = 26 kg/kW, but achieved at a rational cooling capacity q0.15rat less than its oversized maximum value q0.15max = 31 kW/(kg/s) by Δq0.15 = q0.15max − q0.15rat of about 7 kW/(kg/s), i.e., more than 20%. When cooling ambient air to ta2 = 10 °C a maximum value of (∑b–∑bopt)/q0.10 takes place at the rational cooling capacity q0.10rat = 35 kW/(kg/s) and provides annual specific fuel reduction ∑b10rat = 47 kg/kW that is close to its maximum value ∑b10max = 49 kg/kW and achieved at a rational cooling capacity q0.10rat less than its

Energies 2020, 13, 6201 13 of 22

Here: q0.opt—optimal, q0.rat—rational and q0.max—maximum values; points O15 and O10 correspond to q and q ;R and R —for q and q ; ∆q = q q 0.15opt 0.10opt 15 10 0.15rat 0. 10rat 0.15 0.15max − 0.15rat and ∆q0.10 = q0.10max q0.10rat reduction of design specific cooling capacity q0. − − P As Figures9 and 10 show, a maximum rate of annual specific fuel reduction increment ( b P − bopt)/q0.15 when cooling ambient air to ta2 = 15 ◦C takes place at specific cooling capacity q0.15rat = P 24 kW/(kg/s) and provides annual specific fuel reduction b15rat very close to its maximum value P b15max = 27 kg/kW, but achieved at a rational cooling capacity q0.15rat = 24 kW/(kg/s) that is less than its oversized maximum value q0.15max = 31 kW/(kg/s) by ∆q0.15 = q0.15max q0.15rat of about 7 kW/(kg/s), − P P i.e., more than 20%. When cooling ambient air to ta = 10 C a maximum value of ( b bopt)/q 2 ◦ − 0.10 takes place at the rational cooling capacity q0.10rat = 37 kW/(kg/s) and provides annual specific fuel P P reduction b10rat = 47 kg/kW that is close to its maximum value b10max = 48 kg/kW and achieved at a rational cooling capacity q0.10rat = 37 kW/(kg/s) less than its oversized maximum value q0.10max = 44 kW/(kg/s) by ∆q = q q of about 9 kW/(kg/s), i.e., approximately 20% less. The points 0.10 0.10max − 0.10rat R15 and R10 on the TIAC cumulative characteristics of dependence of annual specific fuel reduction P P b15 = f (q0) and b10 = f (q0) on the specific cooling capacities q0.15 and q0.10 are determined according to rational values of design specific cooling capacities q0.15rat and q0.10rat, calculated at the stage 3 of P design methodology. As it is seen, the values of rational annual specific fuel reduction b15rat and P b10rat, determined proceeding from the points R15 and R10, are very close to their maximum values P P b15max and b10max. Graphs in Figures9 and 10 allow users to match current cooling duties and provide minimum sizing system at maximum rate of annular cooling effect increment due to operation at optimal cooling capacity q0.10opt (Figure9) or to peak cooling needs without oversizing due to running at rational cooling capacity q0.rat (Figure 10).

4. Discussion To generalize the results of analysing the efficiency of engine intake air cooling in TIAC systems, using chillers of different types, the optimal and rational design thermal loads for cooling intake air to target temperatures 7 to 15 ◦C in actual site climatic conditions were calculated step-by-step year round on the basis of the proposed novel methodology (Figures 11–14). P As Figure 11 shows, a maximum rate of annual specific fuel reduction b/q0 due to cooling ambient air to the temperatures ta2 = 7, 10 and 15 ◦C takes place at the optimal design specific cooling capacities (points O7,O10,O15): q0.7opt = 33 kW/(kg/s), q0.10opt = 27 kW/(kg/s) and q0.15opt = 16 kW/(kg/s), P P P which provide annual specific fuel reduction b7opt = 56 kg/kW, b10opt = 42 kg/kW and b15opt = P 22 kg/kW considerably less than their maximum values. These values of bopt are used to determine P P P the range of annual specific fuel reduction ( b bopt) beyond their optimal values bopt to calculate P P − the ratios ( b bopt)/q0 that characterize the rate of annual specific fuel reduction beyond the optimal P − values bopt and their maximum values correspond to rational design specific cooling capacities: q0.rat. P P As Figures 11 and 12 show, a maximum rate of annual fuel reduction increment ( b bopt)/q0 in P − the ranges beyond their optimal values bopt (Figure 11) is characterized by maximum values of ratios P P ( b bopt)/q which correspond to rational design specific cooling capacities q when cooling − 0 0.rat ambient air to the temperatures ta2 = 7, 10 and 15 ◦C: q0.7rat = 41 kW/(kg/s), q0.10rat = 35 kW/(kg/s) and q0.15rat = 24 kW/(kg/s) (stage 3, Figure 12). These values of rational design specific cooling capacities q0.rat determine the position of the points R7,R10 and R15 on the TIAC cumulative characteristics P P P of dependence of annual specific fuel reduction b7 = f (q0), b10 = f (q0) and b15 = f (q0) and P P P corresponding rational annual specific fuel reduction b7rat, b10rat and b15rat, calculated at the stage 4 of design methodology (Figure 13). P P P As it is seen, the values of rational annual specific fuel reduction b7rat, b10rat and b15rat, determined proceeding from the points R7,R10 and R15, are very close to their maximum values. Energies 2020, 13, x FOR PEER REVIEW 14 of 23

2st Σb, g/kW Σb/q0,g/kW/(kg/s) 1800 70,00070000 t =7,10,15°C Σb/q0.7 ∑b7 1600 2 Σb/q0.10 O7 ∑b7opt 60,00060000 1400 ∑b Σb/q0.15 10 50,00050000 1200 ∑b10opt O10 1000 40,00040000

800 ∑b15 30,00030000 ∑b15opt 600 O15 20,00020000 400 200 10,00010000 0 0 0 q15opt q10optq7opt 0 5 10 15 20 25 30 35 40 45q 50, kW/(kg/s) 55 0

Figure 11. Relative increment of annual specific fuel reduction ∑b/q0 and annual specific fuel reduction ∑b due to cooling air to ta2 = 7, 10 and 15 °C against a cooling capacity needed q0 (2 stage of methodology).

As Figures 11 and 12 show, a maximum rate of annual fuel reduction increment (∑b − ∑bopt)/q0 in the ranges beyond their optimal values ∑bopt (Figure 11) is characterized by maximum values of ratios (∑b − ∑bopt)/q0 which correspond to rational design specific cooling capacities q0.rat when cooling ambient air to the temperatures ta2 = 7, 10 and 15 °C: q0.7rat = 41 kW/(kg/s), q0.10rat = 35 kW/(kg/s) and q0.15rat = 24 kW/(kg/s) (stage 3, Figure 12). These values of rational design specific cooling capacities q0.rat determine the position of the points R7, R10 and R15 on the TIAC cumulative characteristics of dependence of annual specific fuel reduction ∑b7 = f(q0), ∑b10 = f(q0) and ∑b15 = f(q0) and corresponding Energiesrational2020 annual, 13, 6201 specific fuel reduction ∑b7rat, ∑b10rat and ∑b15rat, calculated at the stage 4 of design14 of 22 methodology (Figure 13).

Σb/q0,10(Σb-Σbopt)/q0, g/kW/(kg/s) 3st 2000 t2=7,10,15°C R 1800 15 O 7 R 1600 7 R 1400 O10 10 O15 1200 Σb/q0.7 Σb/q 1000 0.10 Σb/q 800 0.15 Energies 2020, 13, x FOR PEER600 REVIEW 15 of 23 400 oversized maximum value q0.10max = 44 kW/(kg/s) by Δq0.10 = q0.10max − q0.10rat of about 9 kW/(kg/s), i.e., 200 q7opt approximately 20% less. 0 q 15opt q15rat q10opt q10rat q7rat 0 5 10 15 20 25 30 35 40 45q , kW/(kg/s) 50 55 0 Energies 2020, 13, x FOR PEER REVIEW Σb, g/kW 15 of 23 10(Σb-Σbopt)/q0, g/kW/(kg/s) 4st Figure 12. Relative increments of annual specific fuel reduction ∑b/qP0 and (∑b − ∑Pbopt)/Pq0·due to cooling Figure 12. Relative increments of annual specific fuel reduction b/q0 and70,000 ( b bopt)/q0 due to 2000 ∑b 70000 − · air ta2 = 7, 10 and 15 °C againstt2=7,10,15°C specific cooling capacitiesΣb neededR7 q0 (37rat stage): −∑b/q0; −(∑Pb − ∑boptP)/q0. oversizedcooling maximum air ta2 = 7, value 10 and1800 q 150.10max◦C against= 44 kW/(kg/s) specific cooling by Δ7q capacities0.10 = q0.10max needed − q0.10ratq0 (3 of stage): about 9b kW/(kg/s),/q0; ( b i.e., P 60,00060000 − − − approximatelybopt)/q0. 20% less. 1600 R Σb10 ∑b As it is seen, the 1400values of rational annual specific10 fuel reduction10rat 50,00050000 ∑ b7rat, ∑b10rat and ∑b15rat, determined proceeding 1200from the points R7, R10 and R15, are very close to40,000 their maximum values. Σb, 40000g/kW 100010(Σb-Σbopt)/q0, g/kW/(kg/s) 4st Σb opt 0.15 As Figure 13 shows, a maximum rate of annual15 specific fuel reduction∑b15rat 30,00030000 increment (∑b − ∑b )/q 2000800 ∑b 70,00070000 t2=7,10,15°C Σb R7 7rat when cooling ambient air1800600 to ta2 = 15 °C takesR place at7 specific cooling capacity q0.15rat = 24 kW/(kg/s) 15 Δq0.7 60,0006000020,00020000 and provides annual specific1600 fuel reduction ∑b15rat very closeΔq to its maximum value ∑b15max = 26 kg/kW, 400 Δq 0.10Σb ∑b 0.15 R10 10 10rat 50,0005000010,00010000 1400200 0.15rat 0.15max but achieved at a rational cooling capacity q less than itsq 10maxoversized maximum value q = 31 12000 0 0 kW/(kg/s) by Δq0.15 = q0.15max − q0.15rat of about 7q 15ratkW/(kg/s),q15 max q10rat i.e.,q7rat moreq 7max than40,00040000 20%. When cooling ambient 1000 0 5 10 15 20 25 30 35 40 45q 50, kW/(kg/s) 55 Σb15 0∑b air to ta2 = 10 °C a maximum800 value of (∑b–∑bopt)/q0.10 takes place at the15rat rational30,00030000 cooling capacity q0.10rat = 35 kW/(kg/s) and provides annual specificR fuel reduction ∑b10rat = 47 kg/kW that is close to its Figure 13. Annual specific600 fuel reduction ∑b 15and its relative incrementsΔq0.7 20,00020000(∑b − ∑bopt)/q0·for ta2 = 7, 10 maximum value ∑0 b10max 400= 49 kg/kW and achieved at aΔ qrational0.10 coolingopt capacity0. q0.10rat less than its and 15 °C: Δq —reduction of design specific coolingΔq0.15 capacity;·−(∑b − ∑b 10,000)/q 200 10000 q 10max 0 0 q15rat q15 max q10ratq7rat q 7max 0 The values of rational annu0al 5 specific 10 15 fuel 20 25reduction 30 35 ∑ 40b7rat 45, q∑ 50b, 10ratkW/(kg/s) 55 and ∑b15rat, proceeding from the 0 points R7, R10 and R15, determined according to rational cooling capacities q0.rat, are very close to their

Figure 13. Annual specific fuel reductionP ∑b and its relative increments0.rat P (∑b −P ∑bopt)/q0·for ta2 = 7, 10 maximumFigure values, 13. Annual but specific achieved fuel at reduction rational coolingb and its capacities relative increments q less (thanb theirbopt oversized)/q for ta = maximum7, 10 − 0· 2 0 P Popt 0. valuesandand by 15 15 about °C:C: Δ∆q 20%——reductionreduction (Figure of 14). design specificspecific cooling capacity;capacity;·−((∑bb − ∑bbopt)/q)/q . ◦ 0 ·− − 0

The values of rational Σannub, g/kWal specific fuel reduction ∑b7rat, ∑b10rat and ∑b15rat, proceeding from the 1-4st t2=7,10,15°C points R7, R10 and R15, determined according to rational cooling capacities q0.rat, are very close to their 70,00070000 ∑b R Σb7 maximum values, but achieved at rational7rat cooling capacities q7 0.rat less than their oversized maximum 60,00060000 ∑b O7 values by about 20% (Figure 14). 7opt R 50,00050000 ∑b10rat 10 Σb10 ∑b10opt O10 40,00040000Σb, g/kW 1-4st t2=7,10,15°C 70,0007000030,00030000 ∑b15rat R15 Σb15 ∑b7rat R7 Σb7 ∑b15opt O15 60,0006000020,00020000 ∑b7opt O7 Δq0.7 Δq0.15 R Δq0.10Σb 50,0005000010,00010000 ∑b10rat 10 10 ∑b10opt O10 40,000400000 0 q q 15opt 15ratq10opt q7optq10rat q7rat 0 5 10 15 20R 25 30 35 40 45q0, kW/(kg/s) 50 55 30,00030000 ∑b15rat 15 Σb15 ∑b15opt O15 20,00020000 P 0 a2 FigureFigure 14.14. Annual valuesvalues ofof specificspecific fuelfuel reductionreduction ∑bbagainst againstspecific specificΔq0.7 cooling cooling capacities capacitiesq 0q for fort at2 == 7, Δq0.15 Δq0.10 1010 andand 1515 ◦°C:C: ∆Δqq00—reduction—10,00010000reduction of ofdesign designspecific specific cooling cooling capacity. capacity.

00 q q q q q q P P 15opt 15rat 10opt 7opt 10rat 7rat b b q AsTo Figuregeneralize 13 shows, the results a maximum 0of analysing 5 10 rate 15 of the annual 20 efficien 25 specific 30cy 35 of fuelengine 40 reduction45q , kW/(kg/s)intake 50 55 incrementair cooling ( in TIAC optsystems)/ 0.15 0 − whenby chillers cooling of ambientdifferent air types to ta2 the= 15 TIAC◦C takes cumulative place at specificcharacteristics cooling capacityof dependenceq0.15rat = of24 annual kW/(kg specific/s) and P P providesfuel Figurereduction annual 14. Annual ∑b specific = f(q values0) on fuel the of reduction specificspecific fuel cooling breduction15rat capacitiesvery ∑ closeb against q0 to areits specific presented maximum cooling in value relativecapacitiesb 15maxvalues q0 for= tasa262 =referred kg7, /kW, butto their achieved10 and maximum 15 °C: ata Δ rationalq values,0—reduction coolingselected of design capacity on the specific basisq0.15 ratcoolingof lesstraditional thancapacity. its methods oversized of maximumdesigning, valuei.e., asq 0.15max∑b/bmax= = f(q0/q0.max) (Figure 15). To generalize the results of analysing the efficiency of engine intake air cooling in TIAC systems by chillers of different types the TIAC cumulative characteristics of dependence of annual specific fuel reduction ∑b = f(q0) on the specific cooling capacities q0 are presented in relative values as referred to their maximum values, selected on the basis of traditional methods of designing, i.e., as ∑b/bmax = f(q0/q0.max) (Figure 15).

Energies 2020, 13, 6201 15 of 22

31 kW/(kg/s) by ∆q0.15 = q0.15max q0.15rat of about 7 kW/(kg/s), i.e., more than 20%. When cooling − P P ambient air to ta2 = 10 ◦C a maximum value of ( b– bopt)/q0.10 takes place at the rational cooling P capacity q0.10rat = 35 kW/(kg/s) and provides annual specific fuel reduction b10rat = 47 kg/kW that is P close to its maximum value b10max = 49 kg/kW and achieved at a rational cooling capacity q0.10rat less than its oversized maximum value q = 44 kW/(kg/s) by ∆q = q q of about 0.10max 0.10 0.10max − 0.10rat 9 kW/(kg/s), i.e., approximately 20% less. P P P The values of rational annual specific fuel reduction b7rat, b10rat and b15rat, proceeding from the points R7,R10 and R15, determined according to rational cooling capacities q0.rat, are very close to their maximum values, but achieved at rational cooling capacities q0.rat less than their oversized maximum values by about 20% (Figure 14). To generalize the results of analysing the efficiency of engine intake air cooling in TIAC systems by chillers of different types the TIAC cumulative characteristics of dependence of annual specific fuel P reduction b = f (q0) on the specific cooling capacities q0 are presented in relative values as referred to P their maximum values, selected on the basis of traditional methods of designing, i.e., as b/bmax = f (q0/q0.max) (Figure 15). Energies 2020, 13, x FOR PEER REVIEW 16 of 23

t2=7,10,15°C Σb/Σbmax R15,R10,R7 1 rat15, rat10, rat7 opt7 O 0.9 10 O opt10 O 7 0.8 15 opt15 0.7 t =7°C 0.6 2 t =10°C 0.5 2 t =15°C 0.4 2 0.3 0.2 0.1

0 opt7 rat15 rat7 opt15 opt10 rat10 0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9q /q 1 0 0.max P Figure 15. Relative values of annualannual fuelfuel reductionreduction ∑bb/b/bmaxmax against relativerelative coolingcooling capacities capacitiesq 0q/0q/q0.max0.max needed for cooling ambient air to ta22 == 7,7, 10 10 and and 15 15 °C.◦C.

As shown in Figure 15,15, the application application of developed developed novel methodology allows to determine the precise valuesvalues ofof rationalrational design design specific specific cooling cooling capacities capacitiesq0.rat q0.rat(corresponding (corresponding overall overall values valuesQ0.rat Q0.rat) of) ofTIAC TIAC systems systems that that provide provide annual annual fuel fuel reduction reductio veryn very close close to to their their maximum maximum values, values, but but achieved achieved at atcooling cooling capacities capacities less less than than the values,the values, defined defined by traditional by traditional methods methods issuing issuing from maximumfrom maximum yearly yearlythermal thermal loads or loads maximum or maximum annual annual effect due effect to intakedue to airintake cooling, air cooling, by about by 20%. about 20%. As well as the proposed absorption-ejector TIAC systems, there are the advanced versions of typical basic absorption TIAC system, the the economic comparison, comparison, where where the the latter latter might might be be done done by by taking into account only only the the cost cost of of extra heat exchangers exchangers of of ECh ECh (refrigerant (refrigerant evaporator-air evaporator-air cooler, cooler, refrigerant condenser and ejector) with unchangedunchanged maintenance cost, personal costs, etc.etc. As a basic variant,variant, thethe GTGT of of power power output output of of 10 10 MW MW and and air air mass mass flow flow rate rate Ga G= a40 =40 kg kg/s/s might might be beaccepted accepted for for easy easy recalculation recalculation for for turbines turbines with wi otherth other characteristics. characteristics. According According to Figureto Figure 14 q140.10 qrat0.10rat= =35 35 kW kW/(kg/s)/(kg/s) for for TIAC TIAC system system with with AECh AECh and andq 0.15ratq0.15rat == 24 kW/(kg/s) kW/(kg/s) for traditional ACh or their total values Q0.100.10ratrat == 14001400 kW kW and and QQ0.15rat0.15rat = =960960 kW kW for for air air mass mass flow flow rate rate G Gaa == 4040 kg/s kg/s with with corresponding corresponding P P annual specificspecific fuelfuel reductionreduction ∑bb10rat10rat == 4747 kg/kW kg/kW and ∑bb15rat15rat == 2626 kg/kW kg/kW or or annual annual total total fuel fuel reduction reduction P P ∑BB10rat10rat = =470470 t and t and ∑B15ratB15rat = 260= t.260 Proceeding t. Proceeding from these from thesedata, the data, value the of value the total of the cooling total cooling load on loadlow temperatureon low temperature ECh Q0.10-15 EChrat=Q Q0.10-150.10rat −rat Q=0.10Qrat0.10 i.e.,rat 440Q kW,0.10rat providesi.e., 440 additional kW, provides annual additional fuel reduction annual ∑ fuelB10- P P P − 15ratreduction =∑B10rat −B ∑B15rat =aboutB 210 t asB comparedabout 210 with t as traditional compared TIAC with traditionalsystem with TIAC ACh. system Thus, with the 10-15rat 10rat − 15rat coolingACh. Thus, capacity the cooling of refrigerant capacity evaporator-air of refrigerant evaporator-air cooler of ECh cooler Q0.10-15 ofrat ECh= 440Q 0.10-15kW andrat= 440of refrigerant kW and of condenser Qcondenser = Q0.10-15rat (1 + COP), i.e., about 1760 kW, where COP of ECh can be accepted as 0.3 [42,43]. Taking into account the cost of extra heat exchangers of ECh about 150,000$ according to [75] and including addition 10% increase for ejector and 10% for their mounting the cost of addition equipment of ECh is about 180,000$. On the other hand, the cost of gas fuel annually saved is about $45,000 (proceeding from the price of gas $150 per 1000 m3) and the payback period is about four years. Because of the fluctuations in cost of heat exchangers of different manufacturers and the price of the gas fuel especially the economic analysis is to be conducted for the concrete case. Thus, we use the considered method of designing focuses to provide just initial basic data as rational technical characteristics for further complicated detailed economic analysis. The proposed approach and a novel methodology of rational designing of engine intake air cooling systems enables saving about 20% of financial expenses and to cover actual yearly cooling duties without cooling system oversizing and unproductive decrease of engine power output to overcome the intake air pressure drop in oversized air coolers. Such a method is quite quantitative to evaluate the effect due to cooling and select the rational design cooling capacity and estimate a decrease in sizes of chillers.

Energies 2020, 13, 6201 16 of 22

refrigerant condenser Qcondenser = Q0.10-15rat (1 + COP), i.e., about 1760 kW, where COP of ECh can be accepted as 0.3 [42,43]. Taking into account the cost of extra heat exchangers of ECh about 150,000$ according to [75] and including addition 10% increase for ejector and 10% for their mounting the cost of addition equipment of ECh is about 180,000$. On the other hand, the cost of gas fuel annually saved is about $45,000 (proceeding from the price of gas $150 per 1000 m3) and the payback period is about four years. Because of the fluctuations in cost of heat exchangers of different manufacturers and the price of the gas fuel especially the economic analysis is to be conducted for the concrete case. Thus, we use the considered method of designing focuses to provide just initial basic data as rational technical characteristics for further complicated detailed economic analysis. The proposed approach and a novel methodology of rational designing of engine intake air cooling systems enables saving about 20% of financial expenses and to cover actual yearly cooling duties without cooling system oversizing and unproductive decrease of engine power output to overcome the intake air pressure drop in oversized air coolers. Such a method is quite quantitative to evaluate the effect due to cooling and select the rational design cooling capacity and estimate a decrease in sizes of chillers.

5. Conclusions The efficiency of cooling ambient air at the inlet of gas turbines in temperate climatic conditions by waste heat recovery chillers, using the exhaust gas heat, was analyzed and reserves for its enhancing through deep cooling were revealed. The gas turbine intake air cooling by combined absorption-ejector chillers (AECh) with ACh as a high temperature cooling stage and ECh as a low temperature cooling stage was proposed. It was shown that turbine intake air cooling to about 7 to 10 ◦C in temperate climatic conditions by combined absorption-ejector chillers provides an annual fuel saving of about 50% greater than its value gained due to traditional air cooling to about 15 ◦C in absorption lithium-bromide chiller of a simple cycle. As a result, a novel trend in engine intake air cooling in temperate climatic conditions by two-stage cooling in chillers of combined type has been proposed. A method of logical analysis of the actual operation efficiency of turbine intake air cooling systems in site varying environment, supplemented by the simplest numerical simulation, was used to synthesize new solutions. The real input data on site actual climatic conditions (ambient air temperature tamb and relative humidity ϕamb) were taken by using the well-known program “meteomanz” [74]. The annual fuel reduction was considered to be a primary criterion as an example. The novelty of the research methodological approach to analysis of the TIAC system operation efficiency consists of summation of the current changeable effect (fuel reduction) due to cooling air over the year as an annular effect (annual fuel saving). The rate of its increment according to cooling capacity is used as an indicator to efficient realization of the design cooling capacity. The improved methodology of TIAC system designing, based on this methodological approach, is developed. It involves the stages to determine an optimal value of cooling capacity, providing the minimum system sizes at maximum rate of annual effect increment, and its rational value, providing a close to maximum annual effect without TIAC system oversizing at the second maximum rate of annual effect increment within the range beyond the first maximum rate. Such a method is quite quantitative when evaluating the effect due to cooling, selecting design cooling capacity, and estimating a decrease in sizes of chillers. The rational value of design cooling capacity (thermal load) provides practically maximum annual fuel saving but with reduced by about 15 to 20% sizes of cooling systems due to correspondingly reduced design cooling capacity of the systems as compared with their values defined by traditional designing focused to cover maximum current thermal loads. The method allows users to peak maximum cooling needs without oversizing system or provide minimum sizing TIAC system at maximum rate of annular cooling effect. The last variant of TIAC Energies 2020, 13, 6201 17 of 22 system operation is very reasonable for applying the energy saving technologies, for instance, based on the thermal storage with accumulating excessive (not consumed) cooling capacities at lowered current thermal loads on the cooling system to cover the peak loads.

Author Contributions: Conceptualization, A.R. (40%), M.R. (30%), E.T. (10%), K.K. (10%) and D.M. (10%); methodology, A.R. (30%), E.T. (30%) and M.R. (30%).; software, A.R.; validation, A.R. (30%). and M.R. (30%); formal analysis, A.R. (30%), E.T. (30%), K.K. (30%), D.M. (30%) and M.R.; writing—original draft preparation, A.R. (30%), E.T. (30%) and M.R. (30%); writing—review and editing, D.M. (50%) and K.K. (50%). All authors have read and agreed to the published version of the manuscript. Funding: This research received no external funding. Conflicts of Interest: The authors declare no conflict of interest.

Nomenclature

AC air cooler ACHT high-temperature air cooler ACLT low-temperature air cooler ACh absorption lithium-bromide chiller AECh absorption-ejector chiller CAECh cascade absorption-ejector chiller CDH cooling degree hour; CDH = ∆t τ · COP coefficient of performance ECh ejector chiller GT gas turbine O optimal point for maximum rate of annual fuel reduction increment R rational point for close to maximum annual fuel reduction SAECh stage absorption-ejector chiller SCAECh stage-cascade absorption-ejector chiller TIAC turbine intake air cooling Symbols B total mass fuel consumption decrease, B = CDH (∆be/∆t) Pe. · be specific fuel consumption ca specific heat of humid air CDH CDH = ∆t τ · damb ambient air absolute humidity Ga air mass flow rate Pe power output Q0 total cooling capacity, heat flow rate q0 specific cooling capacity—per unit air mass flow rate t temperature tamb ambient air temperature ta2 outlet air temperature t0 refrigerant boiling temperature ξ specific heat ratio of the total heat (latent and sensible) rejected from air to its sensible heat τ time interval ϕamb ambient air relative humidity ∆be specific fuel consumption decrease ∆t air temperature decrease P B7,10,15 annual total fuel reduction due to cooling turbine intake air to temperatures 7, 10, 15 ◦C P annual specific fuel reduction (per 1 kW turbine power output) due to cooling turbine b7,10,15 intake air to temperatures 7, 10, 15 ◦C Subscripts a air amb ambient max maximum opt optimal rat rational Energies 2020, 13, 6201 18 of 22

Appendix A

Energies 2020, 13, x FOR PEER REVIEW 19 of 23

"Climate" Input: time, region. Program: "mundomanz.com"

Output: ambient air tamb , φamb

"Thermal (cooling) load" (ambient air cooling process)

Input: tamb ; φamb ; Ga ;

ACh: ta2 = 15 °С; AECh: ta2 = 10 °С; Δt = t − t ∙∙∙ a amb a2 "Gas Turbine climatic characteristic" Equations, correlations: Input: cooling process in dI-diagram: GT climatic characteristic tamb ; φamb ; ta2 = 15 °С; tw = 7 °С for ACh; be = f(tamb ), g/(kWh). tamb ; φamb ; ta2 = 10 °С; t0 = 4 °С for AECh; −3 Output: Δbe /Δt =0.35·10 kg/(kWh·K) tamb ; φamb ; ta2 = 7 °С; t0 = 2 °С for cascade AECh;

total Q0 = Ga сa ∙Δta ∙ξ, kW/(kg/s) or kJ/kg;

specific q0 =Q0 /Ga ; q0 = сa ∙Δta ∙ξ, kW/(kg/s) or kJ/kg;

where ξ= q0 /(сa ·Δta )

Output: Q0 = f(tamb , φamb , ta2 ); q0 = f(tamb , φamb , ta2 ).

"Cooling degree-hour (CDH) potential":

Input: tamb ; φamb ;

ACh: ta2 = 15 °С; AECh: ta2 = 10 °С;

Δta = tamb − ta2 ; Δta , Q0 , q0 .∙∙∙ 4 Gas Turbine: Pe = 10 kW; Ga = 40 kg/s, Δbe /Δt . Equations, correlations:

ACh: Δt15 = tamb − ta2 ; ta2 = 288 K (15 °С); AECh: Δt10 = tamb − ta2 ; ta2 = 283 K (10 °С);

current CDR= Δta ∙τ, K·h;

annual ΣCDH =Σ(Δta ∙τ) , K·h. Annual cumulative characteristic:

ΣCDH = f(q0) for ta2 = 7; 10 (AECh); 15 °С (ACh).

Σbe = Σ[(Δbe /Δt ) Δt ∙τ] , kg/kW; ΣΔB = Σ[Pe (Δbe /Δt ) Δt ∙τ] , kg Output: ΣCDH = Σ(Δt ∙τ) , K·h;

Σbe = Σ[(Δbe /Δt ) Δt ∙τ] , kg/kW;

ΣΔB = Σ[Pe (Δbe /Δt ) Δt ∙τ] , kg

"Rational cooling load of TIAC system (air cooler)":

Input: tamb ; φamb ; Δta = tamb – ta2 ;

ACh: ta2 = 15 °С; AECh: ta2 = 7, 10 °С; 4 Gas Turbine: Pe = 10 kW; Ga = 40 kg/s

Q0 = f(tamb , φamb , ta2 ); q0 = f(tamb , φamb , ta2 ). Equations, correlations: Criterion—annual total fuel reduction

ΣΔB = Σ[Pe (Δbe /Δt ) Δt ∙τ] , kg 1-stage: Cumulative characteristic of TIACS:

ΣΔB = f(Q0);

2-stage: indicator—relative value ΣΔB /Q0 Relative characteristic ΣΔB /Q0 = f(Q0);

Output: Q0.opt , ΣΔBopt at [ΣΔB /Q0]max

3-stage: indicator—relative value Σ(ΔB–ΔBopt )/Q0

Σ(ΔB–ΔBopt )/Q0 = f(Q0);

Output: Q0.rat , ΣΔBrat at [Σ(ΔB–ΔBopt )/Q0]max

4-stage: indicators [ΣΔB /Q0]max , [Σ(ΔB–ΔBopt )/Q0]max

Σ(ΔB–ΔBopt )/Q0 = f(Q0); ΣΔB = f(Q0);

Output: Q0.rat , ΣΔBrat at [Σ(ΔB–ΔBopt )/Q0]max

Q0.opt , ΣΔBopt at [ΣΔB /Q0]max

FigureFigure A1. A1.Flow Flow chart of of calculation calculation procedure. procedure.

Energies 2020, 13, 6201 19 of 22

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